JP2009275919A - Automatic transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To support thrust force generated in both by a low bearing load or a low case load, in a multistage automatic transmission using a speed reduction planetary gear and a planetary gear set. <P>SOLUTION: This automatic transmission includes the speed reduction planetary gear G1 connected to an input shaft 11 and outputting speed reduction rotation to an output element C1 by fixing a reaction element S1, and the planetary gear set G2 for outputting shifting rotation with the speed reduction rotation from its planetary gear as input. A transmission passage is arranged for transmitting the thrust force respectively generated in one element R1 of the speed reduction planetary gear and one element S3 of the planetary gear set in at least first speed driving. The twisting direction of a helical gear of the respective elements is set in its passage so that the direction of the thrust force F1 of the element R1 and the direction of the thrust force F3 of the element S3 become mutually the different direction in the first speed driving. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to an automatic transmission, and more particularly to a technique for supporting a thrust force generated in a transmission element by power transmission in the gear train.

  In order to improve fuel efficiency, which is indispensable for energy saving as well as ensuring vehicle drivability, there is a demand for a multi-stage automatic transmission for vehicles. To the 5th speed. In order to realize further multi-stage within a limited vehicle mounting space, it is necessary to further reduce the gear train per shift stage and simplify the mechanism, and it is difficult to configure the gear train. In such circumstances, Patent Document 1 proposes a gear train that achieves six forward speeds and one reverse speed by using a planetary gear set including a minimum number of transmission elements. The gear train according to this proposal is characterized in that a multi-stage shift is achieved by a combination of a speed reduction planetary gear and a planetary gear set that receives a speed reduction rotation from the speed reduction planetary gear and outputs a speed change rotation.

JP-A-4-219553

  The gear train configuration according to the above proposal is rational in principle in terms of the number of shift elements per shift stage and the number of required clutches and brakes, but leaves problems to be improved in practice. ing. In particular, in an automatic transmission for a passenger car, in order to reduce gear noise, it is customary to use helical gears for the sun gear, pinion, and ring gear that constitute the planetary gear. The gears increase the meshing ratio of each gear and reduce gear noise, but on the other hand, there is a problem in that opposite thrust forces are generated in the ring gear and the sun gear.

  No solution is specifically disclosed in the above publication for dealing with such thrust force, but the concept on the extension of the conventional technology used in conventional automatic transmissions is applied to the proposed gear train. If the center support is disposed between the planetary gear for reduction (single planetary gear set) and the planetary gear set (double planetary gear set), the thrust force generated in each planetary gear set by power transmission is separated from the center support. Although it will take a method of receiving it in the case so that it does not interfere with each other, if such a configuration is adopted, the axial dimension of the transmission will increase by the amount of the center support, and even if it is multistage, it will be moderate As a countermeasure for a multi-stage automatic transmission in which an increase in the size of the mechanism cannot be avoided, there is a problem particularly in that the mountability on a vehicle is impaired.

  Accordingly, the present invention provides an automatic transmission that solves the above-described problems and can support the thrust forces acting on the planetary gear for deceleration and the planetary gear set, respectively, without increasing the axial dimension of the transmission. For the purpose.

  In order to achieve the above object, the present invention is directed to a speed reduction planetary gear that outputs a speed reduction rotation to an output element by being connected to an input shaft and fixing a reaction force element, and a speed reduction rotation from the speed reduction planetary gear as inputs. In an automatic transmission that achieves multi-stage shifting with a planetary gear set that outputs a variable speed rotation, at least when driving at the first speed, a thrust force generated by one element of the planetary gear for reduction and one element of the planetary gear set is transmitted. The direction of thrust force of the one element of the planetary gear for reduction in the common transmission path and the direction of thrust force of the one element of the planetary gear set in the common transmission path are different from each other during the first speed drive. The twisting direction of the helical teeth of each element is set so as to be in the direction.

  In the above configuration, the torsional direction is a direction in which the thrust force generated in the one element of the planetary gear for reduction and the one element of the planetary gear set faces each other during the first speed drive, and in the common transmission path, A bearing that commonly receives a thrust force generated in each element, the bearing being a thrust of a difference between a thrust force acting on the one element of the planetary gear set and a thrust force acting on the one element of the speed reduction planetary gear; It is effective to receive power.

  Further, in the above configuration, the twisting direction is a direction in which the thrust force generated in the one element of the planetary gear for reduction and the one element of the planetary gear set is separated from each other during the first speed driving, and the common transmission is performed. In the path, there is a bearing for receiving a thrust force generated in each element, and the bearing independently performs a thrust force acting on the one element of the planetary gear set and a thrust force acting on the one element of the planetary gear for reduction. It can also be received.

  In any of the above configurations, it is effective that the reaction force element of the planetary gear for reduction is fixed to the case, and the thrust force transmitted to the common transmission path is transmitted to the case via the bearing. It is.

  In any of the above-described configurations, it is also effective that the planetary gear set carrier is supported in the axial direction on the common transmission path.

  In any of the above-described configurations, the planetary gear set includes a first sun gear that is one element of the planetary gear set, and a second sun gear that is separate from the first sun gear, and has a thrust force acting on the second sun gear. It is effective that the bearing to be received is arranged on a path different from the common transmission path.

  In any one of the above-described configurations, the planetary gear set includes a first sun gear and a second sun gear that is separate from the first sun gear, and a bearing is disposed between the first and second sun gears. It is also effective to adopt a configuration.

  In any of the above-described configurations, it is effective that the one element of the planetary gear for reduction is a ring gear.

  In any of the above-described configurations, the reaction force element of the speed reduction planetary gear is fixed to the oil pump case, and the thrust force generated in the one element of the speed reduction planetary gear is transmitted to the oil pump case via the bearing. It is also effective to adopt the configuration.

  The reduction planetary gear meshes with a sun gear as the reaction force element, a carrier connected to the one element of the planetary gear set supporting a pinion meshing with the sun gear, and a pinion supported by the carrier, It is effective that the ring gear is one element of the planetary gear for reduction connected to the input shaft via a connecting member, and the bearing is arranged between the sun gear and the connecting member.

  The one element of the planetary gear set is a sun gear, the connecting member is provided between the planetary gear for reduction and the planetary gear set, and a second bearing is provided between the connecting member and the sun gear. It is effective to configure.

  The planetary gear set is any one of a long pinion and a short pinion supported by a carrier and meshing with each other, a first sun gear meshing with the long pinion, a second sun gear meshing with the short pinion, and a long pinion and a short pinion. It is effective to adopt a configuration that is a Ravigneaux type planetary gear set including a ring gear that meshes with the ring gear.

  Further, it is effective that the ring gear is configured to mesh with the short pinion.

  In addition, it is effective that the ring gear is configured to mesh with the long pinion.

  Further, it is effective that the long pinion meshes with the first sun gear at one end thereof, and the ring gear meshes with the other end of the long pinion.

  The input shaft is further connected to the planetary gear set carrier via another connecting member, and a third bearing is disposed between the sun gear of the planetary gear set and the other connecting member, A fourth bearing is arranged between the connecting member and the thrust force of the ring gear of the speed reduction planetary gear during reverse drive is transmitted to the case via the second, third and fourth bearings. It is valid.

  In the first aspect of the present invention, during the first speed driving where the driving force is the largest and the load due to the thrust force is applied, the thrust force of each of the speed reduction planetary gear and the planetary gear set depends on the combination of the helical directions of the transmission elements. Since the directions are different from each other, when this direction is a direction facing each other, the common transmission path has a thrust force of one element of the planetary gear set from one direction, and one element of the planetary gear for reduction from the other direction. And the resultant thrust force is received outside the common transmission path. Further, when the above directions are separated from each other, the thrust force does not act on the common transmission path, and as a result, each thrust force is configured to be independently received outside the common transmission path. . Accordingly, it is possible to prevent the durability of the member receiving the thrust force from being lowered. In addition, since the thrust force generated in each of the speed reduction planetary gear and the planetary gear set is opposed to each other and buffered, it is not necessary to provide a center support for receiving the thrust force, and only the corresponding amount of the transmission The axial dimension can be shortened.

  Next, in the configuration according to claim 2, the thrust force acting on one element of the planetary gear set and the thrust force acting on one element of the speed reduction planetary gear are made to oppose each other, thereby being arranged outside the common transmission path. Since the thrust force applied to the bearing can be reduced, the bearing can be made compact, and the axial dimension of the transmission can be shortened.

  According to the third aspect of the present invention, the thrust force acting on one element of the planetary gear set and the thrust force acting on one element of the speed reduction planetary gear at the time of first speed driving where the driving force is the largest and a load due to the thrust force is applied. Since the thrust force applied to the bearings arranged outside the common transmission path can be reduced, the bearings can be made more compact, and the axial dimensions of the transmission can be reduced. Shortening can be achieved.

  Further, when the reaction force element of the speed reduction planetary gear is not fixed to the case, another thrust bearing is required between the reaction force element of the speed reduction planetary gear fixed by an appropriate fixing means and the case. Therefore, an increase in the axial dimension of the transmission is unavoidable, whereas in the configuration according to claim 4, the sun gear is fixed to the case and integrated, so that the thrust force can be obtained only by the bearing. Therefore, the axial dimension can be further shortened.

  In addition, it is necessary to support at least one end of the carrier of the planetary gear set in the axial direction, but when the support is supported outside the transmission path, the carrier support member is interposed on the transmission path of the thrust force, The number of thrust bearings increases accordingly. On the other hand, in the configuration according to the fifth aspect, since the carrier is supported on the common transmission path, the number of thrust bearings can be reduced.

  According to the sixth aspect of the present invention, the thrust force of the common transmission path is applied to the first sun gear of the planetary gear set, whereas the first sun gear that is transmitted through the common transmission path is applied to the second sun gear. And the thrust force of the planetary gear for deceleration can be prevented from acting. Therefore, the bearing that receives the thrust force generated in the second sun gear can be made compact so as to receive only the thrust force of the second sun gear outside the common transmission path.

  According to the seventh aspect of the present invention, since the thrust force generated in the first sun gear can be transmitted between the first sun gear and the second sun gear via the bearing, the thrust force generated in the first sun gear meshes with it. As in the case of transmission through the carrier of the pinion gear, it is possible to eliminate the need for bearings associated with interposing the carrier in the transmission path for thrust force transmission, and thereby in the common transmission path. It is possible to reduce the number of bearings disposed in the.

  Further, in the configuration according to claim 8, the ring gear and the sun gear of the speed reduction planetary gear face each other with the bearing between them with respect to the thrust force applied to them, so that the ring gear transmits to the case via the bearing. Since the thrust force that is generated is offset by the opposing of the sun gear, it is possible to reduce the load on a member such as a case that receives the thrust force via the bearing and to prevent the durability from being lowered.

  Moreover, in the structure of Claim 9, since the thrust force concerning an oil pump case can be reduced, the fall of durability of an oil pump case can be prevented.

  In the configuration of claim 10, the reduction of the thrust force can prevent an increase in the size of the thrust bearing that is brought into contact with the sun gear of the speed reduction planetary gear, so that the bearing is brought into contact with the tooth bottom of the sun gear. Therefore, it is not necessary to increase the root diameter of the sun gear in order to secure the contact surface, and an increase in the radial dimension of the speed reduction planetary gear can be prevented. Even when the bearing is brought into contact with the tooth end portion of the sun gear, it is not necessary to increase the thickness of the bearing race in order to secure the strength, and accordingly, an increase in the axial dimension of the transmission path can be prevented. it can.

  According to the eleventh aspect of the present invention, when the thrust force of the sun gear of the planetary gear set is greater than the thrust force of the ring gear of the speed reduction planetary gear, the thrust force from the sun gear is transmitted to the first bearing via the second bearing. Then, the thrust force of the ring gear of the speed reduction planetary gear is transmitted to the second bearing in opposition to the thrust force of the sun gear. Accordingly, since the thrust force canceled by the common transmission path is received by the first bearing, the bearing can be configured compactly.

  Furthermore, in the structure of Claim 12, by making a planetary gear set into a Ravigneaux type, the axial direction dimension of a planetary gear set can be shortened by sharing a carrier.

  Further, in the configuration of the thirteenth aspect, since the power transmission is performed on the short pinion side that does not generate the rotational moment, the generation of the rotational moment on the long pinion side can be prevented. The load applied to the support portion can be reduced.

  On the other hand, in the structure of the fourteenth aspect, the radial dimension can be shortened as compared with the case where the ring gear is meshed with the short pinion.

  By the way, when using a Ravigneaux type planetary gear set, when the sun gear and the ring gear mesh with each other facing the long pinion, the pinion is long, so that the thrust direction of the sun gear and the ring gear is set in a facing direction. A thrust force is generated in a direction away from the long pinion that meshes with the sun gear and the ring gear. Further, a separating force is generated between the gears as a force that separates in the meshing of the gears and acts in a direction in which they face each other. As a result, the moment due to the thrust force and the moment due to the separation force act on the long pinion in synergy, and the load acting on the support portion of the long pinion becomes large. On the other hand, in the configuration according to claim 15, the torsion angle of the sun gear is set so that the direction of the thrust force generated in the sun gear of the planetary gear set faces the direction of the thrust force generated in the ring gear of the speed reduction planetary gear. As a result, a thrust force different from that of the sun gear is generated in the ring gear of the planetary gear set, and this force acts in a direction to cancel the moment. Therefore, it is possible to reduce the load applied to the support portion of the long pinion.

  Further, in the configuration of the sixteenth aspect, the thrust force generated in the ring gear of the speed reduction planetary gear is received by the case via the second, third and fourth bearings even during the reverse drive. Compared with the method of dividing the thrust force via the clearance, the clearance between the members is reduced, so that the axial dimension can be reduced accordingly.

1 is a skeleton diagram showing a first embodiment of an automatic transmission to which the present invention is applied. It is an operation | movement diagram of the gear train of the said automatic transmission. It is a speed diagram of the gear train. It is a schematic cross section which shows the said gear train more concretely. It is a schematic cross section which shows the thrust force at the time of the driving force transmission of the said gear train about 1st-4th speed for every gear stage. It is a schematic cross section which shows the said thrust force about 5th, 6th speed, and reverse. It is a chart showing the thrust force applied to each bearing of the gear train as a coefficient. It is a schematic cross section of the gear train of the second embodiment. It is a schematic cross section of the gear train of the third embodiment. It is a schematic cross section of the gear train of the fourth embodiment. It is a schematic cross section which shows the thrust force at the time of the 1st speed drive in 4th Embodiment. It is a schematic cross section of the gear train of the fifth embodiment. It is a schematic cross section which shows the thrust force at the time of the driving force transmission of the gear train of 5th Embodiment about 1st-4th speed for every gear stage. It is a schematic cross section which shows the said thrust force about 5th, 6th speed, and reverse. It is a chart showing the thrust force applied to each bearing of the gear train as a coefficient. It is a schematic cross section of the gear train of the sixth embodiment. It is a skeleton figure which shows 7th Embodiment of the automatic transmission to which this invention is applied. FIG. 3 is a layout diagram illustrating an actual shaft positional relationship of the automatic transmission. It is an operation | movement diagram of the gear train of 7th Embodiment. It is a schematic cross section which shows more specifically the gear train of 7th Embodiment. It is a schematic cross section which shows the thrust force at the time of the driving force transmission of the gear train of 7th Embodiment about 1st-4th speed for every gear stage. It is a schematic cross section which shows the said thrust force about 5th, 6th speed, and reverse. It is a chart showing the thrust force applied to each bearing of the gear train. It is a schematic cross section of the gear train of the eighth embodiment. It is a schematic cross section of the gear train of the ninth embodiment.

  Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 is a skeleton view showing a first embodiment in which the present invention is applied to a vertical automatic transmission for a front engine / rear drive (FR) vehicle. This automatic transmission is connected to an input shaft 11 and fixed to a sun gear S1 as a reaction force element to thereby output a reduction rotation to a carrier C1 as an output element, and a speed reduction from the speed reduction planetary gear G1. The planetary gear set G2 that receives rotation and outputs variable speed rotation is an automatic transmission that achieves multi-stage forward 6-speed and reverse-first speed.

  Further, each part will be described in detail. In this automatic transmission, a torque converter 2 with a lock-up clutch 20 connected to an engine (not shown) is disposed at the forefront of the mechanism, and a transmission mechanism is disposed at the rear thereof. Has been adopted. The torque converter 2 includes a pump impeller 21, a turbine runner 22, a stator 23 disposed therebetween, a one-way clutch 24 that rotationally engages the stator 23 with the transmission case 10, and an inner race of the one-way clutch. And a stator shaft 25 for fixing the motor to the transmission case 10.

  The speed reduction planetary gear G1 is constituted by a simple planetary gear, and a ring gear R1 as an input element as one element thereof is connected to the input shaft 11, and a carrier C1 as an output element for speed reduction rotation is connected via a multi-plate clutch C-1. The transmission is connected to the small-diameter sun gear S3 of the planetary gear set G2 and is also connected to the large-diameter sun gear S2 of the planetary gear set G2 through the multi-plate clutch C-3. It is fixed to the case 10.

  The planetary gear set G2 that forms the main body of the speed change mechanism is engaged with a pair of sun gears S2 and S3 having different large and small diameters, one of which meshes with the large diameter sun gear S2 and the other with the ring gear R3, and the other with the small diameter sun gear S3. And a Ravigneaux type gear set comprising a carrier C2 (C3) supporting a long pinion P2 and a short pinion P3. The large-diameter sun gear S2 can be locked to the case 10 by a brake B-1 composed of a band brake, a one-way clutch F-1 in parallel therewith, and a multi-plate brake B-2 that enables the engagement thereof. The carrier C2 (C3) can be locked to the case 10 by a parallel one-way clutch F-2 and a multi-plate brake B-3. Carrier C2 (C3) as an input element for non-decelerated rotation of planetary gear set G2 is connected to input shaft 11 via multi-plate clutch C-2, and ring gear R3 as an output element for variable speed rotation is connected to output shaft 19. It is connected.

  The automatic transmission having such a configuration shifts based on the vehicle load and the vehicle speed within a shift range corresponding to the range selected by the driver under the control of an electronic control device and a hydraulic control device (not shown). Fig. 2 shows the engagement and release of each clutch, brake and one-way clutch (circled for engagement, unmarked for release, △ for engagement only during engine braking, and ● for engagement not directly affecting the achievement of the gear stage. The shift speed achieved in (1) is shown in a graph. FIG. 3 is a speed line showing the relationship between the shift speed achieved by engagement of each clutch, brake, and one-way clutch (represented by the mark ●) and the rotation speed ratio of each shift element at that time. Shown in the figure.

  As can be seen with reference to FIGS. 2 and 3, the first speed (1st) is the engagement of the clutch C-1 and the brake B-3 (in this embodiment, as can be seen with reference to the operation table, The automatic engagement of the one-way clutch F-2 is used instead of the engagement of the brake B-3. The reason why this engagement is used and the reason why this engagement corresponds to the engagement of the brake B-3 Will be described in detail later). In this case, referring to FIG. 1, the rotation reduced from the input shaft 11 via the speed reduction planetary gear G1 is input to the small-diameter sun gear S3 of the planetary gear set G2 via the clutch C-1 and the engagement of the one-way clutch F-2. The reaction force is applied to the carrier C <b> 2 (C <b> 3) that is locked by the combination, and the reduction rotation of the maximum reduction ratio of the ring gear R <b> 3 is output to the output shaft 19.

  Next, the second speed (2nd) is applied to the engagement of the one-way clutch F-1 corresponding to the engagement of the clutch C-1 and the brake B-1 and the engagement of the brake B-2 (these). The reason why the engagement corresponds to the engagement of the brake B-1 will also be described later.) In this case, the rotation decelerated from the input shaft 11 via the deceleration planetary gear G1 is input to the small-diameter sun gear S3 of the planetary gear set G2 via the clutch C-1, and the brake B-2 and the one-way clutch F-1 are engaged. The reaction force is applied to the locked large-diameter sun gear S2, and the reduced rotation of the ring gear R3 is output to the output shaft 19. The reduction ratio at this time is smaller than the first speed (1st) as shown in FIG.

  The third speed (3rd) is achieved by simultaneous engagement of the clutch C-1 and the clutch C-3. In this case, the rotation decelerated from the input shaft 11 via the deceleration planetary gear G1 is simultaneously input to the large-diameter sun gear S2 and the small-diameter sun gear S3 of the planetary gear set G2 via the clutch C-1 and the clutch C-3. Since G2 is in a directly connected state, the same rotation of the ring gear R3 as the input rotation to the sun gears S2 and S3 is output to the output shaft 19 as a rotation decelerated with respect to the rotation of the input shaft 11.

  Further, the fourth speed (4th) is achieved by simultaneous engagement of the clutch C-1 and the clutch C-2. In this case, the rotation decelerated from the input shaft 11 via the deceleration planetary gear G1 is input to the small-diameter sun gear S3 of the planetary gear set G2 via the clutch C-1, and on the other hand from the input shaft 11 via the clutch C-2. The input non-decelerated rotation is input to the carrier C2 (C3), and an intermediate rotation between the two input rotations is output to the output shaft 19 as the rotation of the ring gear R3 slightly decelerated with respect to the rotation of the input shaft 11. Is done.

  Next, the fifth speed (5th) is achieved by simultaneous engagement of the clutch C-2 and the clutch C-3. In this case, the rotation decelerated from the input shaft 11 via the deceleration planetary gear G1 is input to the large-diameter sun gear S2 of the planetary gear set G2 via the clutch C-3, and on the other hand, from the input shaft 11 via the clutch C-2. The non-decelerated rotation input in (1) is input to the carrier C2 (C3), and the rotation slightly increased from the rotation of the input shaft 11 of the ring gear R3 is output to the output shaft 19.

  The sixth speed (6th) is achieved by engagement of the clutch C-2 and the brake B-1. In this case, non-decelerated rotation is input from the input shaft 11 via the clutch C-2 only to the carrier C2 (C3) of the planetary gear set G2, and is opposed to the large-diameter sun gear S2 locked by the engagement of the brake B-1. Then, the rotation of the ring gear R3 further increased is output to the output shaft 19.

  Note that reverse (Rev) is achieved by engagement of the clutch C-3 and the brake B-3. In this case, the rotation decelerated from the input shaft 11 via the deceleration planetary gear G1 is input to the large-diameter sun gear S2 of the planetary gear set G2 via the clutch C-3, and the carrier is locked by the engagement of the brake B-3. The reaction force is applied to C2 (C3), and the reverse rotation of the ring gear R3 is output to the output shaft 19.

  Each speed stage thus achieved is represented by a speed ratio of the ring gear R3 (this ring gear theoretically constitutes a separate ring gear R2 for the large-diameter sun gear S2) on the speed diagram of FIG. As can be seen qualitatively by referring to the vertical intervals of the circles that indicate, the speed steps are relatively good at regular intervals for each gear.

  Here, the relationship between the one-way clutch F-2 and the brake B-3 and the relationship between the one-way clutch F-1 and both the brakes B-1 and B-2 will be described. As can be seen from the engagement / release relationship of the brakes B-1 and B-3 at the first speed and the second speed described above, one of these brakes is released at the time of up-downshifting between both gears. At the same time, the other engagement is performed, which is a so-called gripping friction element. Such gripping of friction elements requires precise simultaneous control of the engagement pressure and release pressure of the hydraulic servo that operates them. To perform such control, the addition of a control valve and complicating the hydraulic circuit, etc. Will be invited. Therefore, in this embodiment, the direction of engagement of the one-way clutch F-2 at the first speed is changed using the reverse reaction torque applied to the carrier C2 (C3) between the first speed and the second speed. By setting it to match the reaction torque support direction, the one-way clutch F-2 performs substantially the same function as the engagement of the brake B-3, and the brake B-3 is engaged at the first speed. (However, since the direction of the reaction torque applied to the carrier C2 (C3) in the wheel-driven vehicle coast state is reversed with respect to the engine-driven state, in order to obtain the engine braking effect, FIG. As shown by the mark, it is necessary to engage the brake B-3) and the carrier C2 (C3) is locked. Therefore, in order to achieve the gear position, it is possible to adopt a configuration in which the first speed is achieved by engaging the brake B-3 without providing a one-way clutch.

  The same relationship as described above holds for the case of the large-diameter sun gear S2, and in this case, the one-way clutch F-1 is set to the one-way clutch F-1 in accordance with the reaction torque support direction at the second speed. The clutch F-1 can exhibit substantially the same function as the engagement of the brake B-1. However, unlike the carrier C2 (C3), the large-diameter sun gear S2 is engaged not only for obtaining the engine braking effect at the second speed but also for achieving the sixth speed. Since it is a speed change element, the brake B-1 is required. Further, as can be seen from the speed diagram of FIG. 3, the large-diameter sun gear S2 rotates in the direction opposite to the input rotation direction when the first speed is achieved. It rotates in the same direction as the rotation direction. Therefore, since the one-way clutch F-1 cannot be directly coupled to the fixed member, the effectiveness of the engaged state can be controlled by the serial arrangement with the brake B-2.

  Next, FIG. 4 is a schematic cross-sectional view that further embodies the transmission mechanism of the gear train. In accordance with the basic feature of the present invention, the automatic transmission transmits the thrust forces F1 and F3 generated in the ring gear R1 of the reduction planetary gear G1 and the small-diameter sun gear S3 of the planetary gear set G2 at least during the first speed drive. A transmission path to be transmitted. The direction of the thrust force F1 of the ring gear R1 of the speed reduction planetary gear G1 and the direction of the thrust force F3 of the small-diameter sun gear S3 of the planetary gear set G2 in this transmission path are different from each other during the first speed drive. The twisting direction of the helical teeth of each element is set. In this embodiment, the direction in which the thrust forces generated in the ring gear R1 of the planetary gear G1 for reduction and the small-diameter sun gear S3 of the planetary gear set G2 face each other during the first speed drive is selected from the different directions. Has been. Specifically, when the rotation direction of the input shaft 11 is viewed clockwise from the front side as shown in the figure, the torsion direction of the helical teeth is the same clockwise for the pinion P1 of the planetary gear G1 for reduction. For the direction and the long pinion P2 of the planetary gear set G2, the torsion in the counterclockwise direction is set. Naturally, the torsional direction of the helical teeth of other elements, including the ring gear R1 meshing with the pinion P1 and the ring gear R3 meshing with the long pinion P2 via the short pinion P3, is a direction commensurate with this.

  This gear train has a large number of bearings in a transmission path for transmitting a thrust force generated in the planetary gear G1 for reduction and the planetary gear set G2. Among these bearings, the first bearing 31 is disposed between the sun gear S1 of the planetary gear G1 for reduction and the connecting member 12 that transmits input rotation to the ring gear R1. The pair of second bearings 32 and 33 are disposed between the connecting member 12 and the extended portion of the small-diameter sun gear S3 of the planetary gear set G2. Further, the pair of third bearings 36 and 37 are disposed between the small-diameter sun gear S3 of the planetary gear set G2 and the other connecting member 13 that transmits the input rotation to the carrier C2 (C3). Similarly, the pair of fourth bearings 38 and 39 are disposed between the other connecting member 13 and the case 10. Accordingly, among these bearings, the first bearing 31 is applied to the thrust force F3 acting on the small-diameter sun gear S3 that is one element of the planetary gear set G2 and the ring gear R1 that is one element of the speed reduction planetary gear G1 in a common transmission path. A thrust force that is a difference from the acting thrust force F1 is received. Further, a pair of bearings 34 receiving thrust forces in both directions of the large-diameter sun gear S2 are provided between the extended portion of the small-diameter sun gear S3 of the planetary gear set G2, the hub side member of the clutch C-1, and the one-way clutch F-2. , 35 are arranged.

  In this embodiment, the sun gear S1 that is a reaction force element of the speed reduction planetary gear G1 is fixed to the case 10. This configuration eliminates the necessity of disposing another thrust bearing between the sun gear and the case as in the case where the sun gear S1 is fixed to the case via other fixing means, and the transmission for that amount. This avoids an increase in the axial dimension. With this configuration, the thrust force transmitted to the common transmission path is transmitted to the case 10 via the bearing 31.

  In this embodiment, the planetary gear set G2 is a Ravigneaux type planetary gear set, the ring gear R3 meshes with the long pinion P2, and the long pinion P2 meshes with the first sun gear S2 having a large diameter at one end thereof. The ring gear R3 is configured to mesh with the other end of the long pinion P2. This configuration is useful for shortening the axial dimension of the planetary gear set G2 by sharing the carrier C2 (C3).

  The input shaft 11 is connected to the carrier C3 (C2) of the planetary gear set G2 via the other connecting member 13 and the clutch C-2, and between the small diameter sun gear S3 of the planetary gear set G2 and the other connecting member 13. The third bearings 36 and 37 are disposed, and the fourth bearings 38 and 39 are disposed between the case 10 and the other connecting member 13. Accordingly, during reverse drive, the thrust force F1 of the ring gear R1 of the speed reduction planetary gear G1 is transmitted to the case 10 via the second bearings 32, 33, the third bearings 36, 37, and the fourth bearings 38, 39.

  5 and 6 are schematic cross-sectional views showing changes in the thrust force at the respective speeds described above. Referring to FIG. 5 (refer to FIG. 4 for the symbols representing the respective members), during the first speed (1st) drive, power transmission is performed by the ring gear R1, the pinion P1 and the carrier C1 of the planetary gear G1 for reduction. The clutch C-1 (the engagement state is indicated by a circle in the figure, the same applies to other engagement elements), the small-diameter sun gear S3, the short pinion P3 and the long pinion P2, and the ring gear R3 of the planetary gear set G2. After that. Therefore, due to the relation of the helical direction of the helical tooth, the forward thrust force (indicated by the left arrow in the figure) generated in the sun gear S1 is supported by the case 10 for the ring gear R1 of the speed reduction planetary gear G1. A backward thrust force F1 (indicated by a rightward arrow in the figure) as a reaction force is transmitted to the connecting member 12 that connects the ring gear R1 to the input shaft 11 as indicated by a thick line in the figure. On the other hand, for the small-diameter sun gear S3 of the planetary gear set G2, the thrust force to the rear of the ring gear R3 (indicated by a right-pointing arrow in the figure) is passed through the bearing 39 as shown by the mark ● in the figure. An extension portion of the sun gear S3 is supported by the latest case 10 and a forward thrust force (indicated by a left-pointing arrow in the figure) F3 as a reaction force is similarly marked with a mark in the figure to indicate a path. To the connection member 12 through the second bearings 33 and 32. In this way, both thrust forces are applied to the connecting member 12 in a direction in which they cancel each other. As a result, a reduced forward thrust force F3-F1 is applied to the bearing 31. Here, the driving at the time of the first speed driving means a state where the transmission is rotated by the engine and the vehicle is going to increase the speed. On the contrary, the coast means a state in which the transmission is rotated by the vehicle (specifically, wheels) and the vehicle is about to decelerate.

  FIG. 7 is a graph showing the thrust force applied to each bearing during driving at each gear stage. As shown in this chart, the thrust force F3 of the planetary gear set G2 is directly applied to the bearings 32 and 33, whereas the thrust force F3 obtained by subtracting the thrust force F1 of the speed reduction planetary gear G1 from the thrust force is applied to the bearing 31. -It can be seen that F1 is applied. The numerical values in this chart represent the thrust force coefficient when the helical angle of the helical tooth is 25 ° for both the reduction planetary gear G1 and the planetary gear set G2. This value is obtained from the torques shared by the large-diameter sun gear S2, the small-diameter sun gear S3, and the ring gear R3 (R2) of the reduction planetary gear G1 and the planetary gear set G2.

  Next, during the second speed (2nd) drive, the power transmission is performed through the small-diameter sun gear S3, the short pinion P3 and the long pinion P2, and the ring gear R3 of the planetary gear set G2, as in the first speed. Is a lock by engagement of the one-way clutch F-1 and the brake B-2, and the large-diameter sun gear S2 also shares the reaction torque, so that the thrust force F2 acts, so The thrust force F4 of the ring gear R3 and the thrust force F2 of the sun gear S2 are applied to the bearing 39, as indicated by the paths marked with and ■. In addition, the thrust force applied to the bearing 31 with a mark ● indicating the path is the same as that in the first speed as shown in FIG. 7.

  Next, at the time of the third speed (3rd) drive, the power transmission is different from the second speed only in that the large-diameter sun gear S2 is rotating, and particularly in terms of torque sharing of elements involved in torque transmission. Since there is no difference, the relationship of the thrust force is exactly the same as that at the time of the second speed drive except that the value becomes smaller as the speed reduction ratio decreases, as will be understood with reference to FIG. .

  Further, during the fourth speed (4th) drive, power transmission is performed in a state where torque transmission from the large-diameter sun gear S2 is lost with respect to the third speed. Therefore, the relationship of the thrust force transmitted forward through the common transmission path marked with ● in the figure is the same as in the case of the third speed drive except that the value is smaller by the reduction of the reduction ratio. It is. Further, the thrust force transmitted rearward has a relationship in which the thrust force F2 of the large-diameter sun gear S2 has been eliminated.

  Next, during the fifth speed (5th) drive shown in FIG. 6, the manner of power transmission is different, and decelerated rotation is input to the large-diameter sun gear S2 of the planetary gear set G2, and non-decelerated rotation is input to the carrier C2 (C3). As a result, the large-diameter sun gear S2 receives a driving torque via the long pinion P2 with respect to the output of the ring gear R3. As a result, the thrust force F2 applied to the large-diameter sun gear S2 is reversed, and this thrust force F2 is applied to the first bearing 31 via the bearing 34 and the respective bearings 33 and 32 of the common transmission path in the path indicated by ●. It takes. And since the thrust force F1 of the same direction by the reduction planetary gear G1 also acts on the 1st bearing 31, the thrust force F1 + F2 which added both is applied. However, at this gear position, the transmission torque is small due to the speed increase, and as is apparent with reference to FIG. 7, the thrust force F2 itself is extremely small compared to the case of the first speed and the second speed. Are smaller than those at the gear positions, and the overlap of thrust forces in the first bearing 31 is not particularly problematic.

  Furthermore, during the sixth speed (6th) drive, power transmission is performed only between the planetary gear set G2 side and between the long pinion P2 and the ring gear R3, and the large-diameter sun gear S2 supports the reaction torque. Become a relationship. At this time, the torque applied to the ring gear R3 and the large-diameter sun gear S2 is further reduced by increasing the speed with respect to the input rotation, so that the thrust force is also reduced. In this case, the thrust force F4 of the ring gear R3 is supported by the case 10 via the bearing 39 along the path indicated by ●, and the thrust force F2 of the large-diameter sun gear S2 is supported by the three bearings 34, 33, 32, the connecting member 12 and the first bearing 31 are supported by the case 10.

  On the other hand, during reverse (Rev) driving, power is transmitted via the long pinion P2 between the large-diameter sun gear S2 of the planetary gear set G2 and the ring gear R3 via the speed reduction planetary gear G1. In this case, since the output of the ring gear R3 is reversely rotated with respect to the input of the large-diameter sun gear S2, the thrust force F4 of the ring gear R3 is offset in a relationship opposed to the thrust force F2 of the large-diameter sun gear S2. Only the thrust force F1 of the ring gear R1 of the planetary gear G1 is transmitted to the case 10 through the bearings 32, 33, 36 to 39 after the second bearing of the common transmission path. However, since both the bearings 37 and 38 in the middle are in the path loop shown in the figure where the thrust force F4 and the thrust force F2 are offset, the thrust force F2 and the thrust force F1 are added. become.

  By the way, since the thrust force at each of the shift speeds is reversed between driving and coasting, the relationship of the thrust force is not established at the time of coasting. However, generally, the transmission torque at the time of coasting is 1/3 of driving. Since the resulting thrust force is extremely small, when viewed from the viewpoint of the bearing load, this is a small value that is almost negligible as compared with driving. Therefore, in order to ensure the durability of the bearing, the relationship of the thrust force during driving is important as described above.

  Further, in the above embodiment, among the bearings, the emphasis is particularly on the reduction of the thrust force applied to the first bearing 31. The bearing rotates at the same speed as the fixed sun gear S1 and the input rotational speed. This is because the bearing load is the largest because it is disposed between the member 12 and the relative rotational speed is large. Incidentally, in the simple comparison of only the thrust force shown in FIG. 7, as a result of the reduction of the thrust force of the first bearing 31 by the application of the present invention, the second bearings 32 and 33 and the fourth bearing 39 are rather at the low speed stage side. Although the applied thrust force is greater, the second bearings 32 and 33 are disposed between the rotating members that originally have a small relative rotational speed difference, and the fourth bearing 39 includes the output member that rotates at a reduced speed by the speed change, and the case 10. Therefore, the endurance load of the bearing that is formed by both the thrust force and the rotational speed is smaller than the load of the first bearing 31.

  As described above, according to the configuration of the first embodiment, the thrust forces F1 and F3 of the deceleration planetary gear G1 and the planetary gear set G2 face each other at the time of first speed driving with the largest driving force and a load caused by the thrust force. Since the direction is set, the thrust force F3 of the sun gear S3 of the planetary gear set G2 acts on the common transmission path from one direction, and the thrust force F1 of the ring gear R1 of the planetary gear G1 for reduction acts from the other direction. As a result, the subtracted thrust force F3-F1 is received outside the common transmission path. Accordingly, it is possible to prevent a decrease in the durability of the bearing 31 that receives the thrust force. Further, since the thrust forces F1 and F3 generated in the planetary gear G1 for reduction and the planetary gear set G2 are opposed to each other and buffer each other, it is not necessary to provide a center support for receiving the thrust force, and correspondingly As much as possible, the axial dimension of the transmission can be reduced.

  Furthermore, since the thrust force applied to the bearing 31 arranged outside the common transmission path can be reduced by making the thrust forces counter to each other as described above, the bearing 31 can be configured in a compact manner. However, the axial dimension of the transmission can be shortened.

  In addition, the reduction of the thrust force can prevent the thrust bearing 31 that contacts the sun gear S1 of the speed reduction planetary gear G1 from being increased in size, and therefore the configuration in which the bearing 31 is brought into contact with the tooth bottom of the sun gear S1 is employed. Therefore, it is not necessary to increase the root diameter of the sun gear S1 in order to secure the contact surface, and an increase in the radial dimension of the speed reduction planetary gear G1 can be prevented. Even when the bearing 31 is brought into contact with the tooth end portion of the sun gear S1, it is not necessary to increase the thickness of the bearing race in order to ensure the strength. Therefore, an increase in the axial dimension of the transmission path is prevented accordingly. be able to.

  Further, the connecting member 12 is disposed between the speed reduction planetary gear G1 and the planetary gear set G2, and the second bearings 32 and 33 are disposed between the connecting member 12 and the sun gear S3. Thrust force F3 of the sun gear S3 of the planetary gear set G2 that is greater than the thrust force F1 of the ring gear R1 of the speed reduction planetary gear G1 is transmitted to the first bearing 31 via the second bearings 33 and 32, and the ring gear R1 of the speed reduction planetary gear G1 The thrust force F1 is transmitted to the second bearings 32 and 33 in opposition to the thrust force F3 of the sun gear S3. Therefore, since the thrust force F3-F1 canceled by the common transmission path is received by the first bearing 31, the bearing can be configured compactly.

  Further, the long pinion P2 of the Ravigneaux type planetary gear set G2 meshes with the first sun gear S2 at one end thereof, and the ring gear R3 is generated at the sun gear S3 of the planetary gear set G2 due to meshing with the other end of the long pinion P2. Since the twist angle of the sun gear S3 is set so that the direction of the thrust force F3 faces the direction of the thrust force F1 generated in the ring gear R1 of the speed reduction planetary gear G1, as a result, the ring gear R3 of the planetary gear set G2 has A thrust force F4 different from that of the sun gear S3 is generated. As a result, the thrust force acting on the sun gear S2 is separated from the ring gear R3. Therefore, the load applied to the support portion of the long pinion P2 can be reduced.

  By the way, in the first embodiment, regarding the configuration of the planetary gear set G2, the ring gear R3 is arranged to mesh with the long pinion P2 on the outer peripheral side of the small-diameter sun gear S3, but the ring gear R3 is arranged on the outer peripheral side of the large-diameter sun gear S2. It is also possible to adopt an arrangement that meshes with the long pinion P2. FIG. 8 is a schematic cross-sectional view of the gear train according to the second embodiment having such an arrangement. Even if such a configuration is adopted, the same thrust force relationship as that of the first embodiment is naturally established, and the same effect can be obtained.

  Next, FIG. 9 shows the configuration of the planetary gear set G2 by reversing the size relationship between the diameters of the first sun gear and the second sun gear with respect to the second embodiment, and causing the short pinion P2 to mesh with the small diameter sun gear S2 and the ring gear R2. A third embodiment in which a long pinion P3 is meshed with a large-diameter sun gear S3 is shown in a schematic cross section. Also in this case, since the generation of the rotational moment in the long pinion P3 can be prevented, the load on the rotation support portion can be reduced. Even if such a configuration is adopted, naturally, the same thrust force relationship as in the first and second embodiments is established, and the same effect can be obtained.

  In each of the above embodiments, the planetary gear set G2 is a Ravigneaux type, but the basic concept of the present invention can also be applied to a planetary gear set in which two ordinary planetary gears are combined. As such an example, finally, a fourth embodiment in which a simple planetary gear G2a and a double planetary gear G2b are combined in place of the Ravigneaux type is schematically shown in FIG.

  In this embodiment, as in the case of using a Ravigneaux type planetary gear set, the large-diameter sun gear S2 and the small-diameter of both planetary gears G2a and G2b are used to improve the speed ratio and speed step obtained for each gear. The sun gear S3 is connected as an input element for reduction rotation via the clutch C-1, and the ring gear R2 of the simple planetary gear G2a is used as an input element for reduction rotation via the clutch C-3. For the double planetary gear G2b, the same helical tooth twist as in each of the above embodiments is set on the pinion P3a on the side meshing with the ring gear R3. In the case of this configuration, the ring gear R2 is connected to the clutch C-3 so as to be an input element for decelerating rotation. Therefore, the connecting portion of the carrier C2 of the pinion P2 and the carrier C3 of the double pinions P3a and P3b is connected to the brake B-3 and It becomes the structure connected with the one-way clutch F-2, and the end of the carrier C2 is supported by the extension part of both sun gears. In this relationship, a bearing 40 is further added.

  When such planetary gear set G2 is used, since both sun gears S2 and S3 are connected, the bearings 32, 33, 34, 35, and 40 are arranged in a common transmission path. The relationship of the thrust force at the time of driving is the same as in the above embodiments. Incidentally, FIG. 11 shows the thrust force at the first speed at which the bearing load becomes the largest, and the pinion meshing with the helical direction of the helical pinion P1 of the reduction planetary gear G1 and the ring gear R3 side of the double planetary gear G2b. By setting the twisting direction of the helical teeth of P3a to the same direction as in the above embodiments, the same thrust force can be supported as shown by the thick dotted line in the figure.

  Each of the above-described embodiments has been aimed at mainly reducing the load applied to the bearing by selecting the facing direction from the concept of making the helical twisting directions different from each other, which is the basis of the present invention. However, by making the different directions different from each other, it can be used mainly for reducing the load transmitted to the case. Next, an embodiment based on this concept will be described.

  FIG. 12 is a schematic cross section showing the configuration of the gear train according to the fifth embodiment. In this embodiment, the rotation direction of the input shaft 11 is clockwise when viewed from the front side, and is opposite to the torsional direction of the helical teeth of FIG. 4 showing the first embodiment, and the pinion of the planetary gear G1 for reduction is used. P1 is twisted counterclockwise and the long pinion P2 of the planetary gear set G2 is twisted clockwise. As a matter of course, the direction of twisting the helical teeth of other elements meshing with these elements is a direction corresponding to this.

  FIGS. 13 and 14 show, on a schematic cross section, changes in thrust force at each gear position when the torsion of the above relationship is set. Referring to FIG. 13 (refer to FIG. 12 for the symbols indicating the respective members), at the time of the first speed (1st) drive, the speed reduction planetary gear G1 A forward thrust force F1 (indicated by a leftward arrow in the figure) F1 acts on the ring gear R1, whereas a rearward thrust force equivalent to that of the sun gear S1 (indicated by a rightward arrow in the figure). However, since these thrust forces are balanced by being transmitted to each other via the connecting member 12 and the first bearing 31 in the transmission path, they constitute other bearings in the transmission path and the front wall of the case. No thrust force acts on the oil pump case 10p. On the other hand, in the planetary gear set G2, a thrust force F3 (represented by a leftward arrow in the figure) F3 acts on the small-diameter second sun gear S3, and a forward thrust force equivalent to the thrust force (shown in the figure) is applied to the ring gear R3. These thrust forces are also applied to each other via the connecting member 14 between the ring gear R3 and the output shaft 19, the third bearings 36 and 37 in the transmission path, and one of the fourth bearings 38. Since it balances by being transmitted, it does not catch on the other bearing of a transmission path. Accordingly, during the first speed (1st) drive, neither the thrust force generated in the planetary gear G1 for reduction nor the planetary gear G2 is transmitted to the case 10.

  FIG. 15 graphically shows the thrust force applied to each bearing during driving at each gear position (the numerical values in the chart are obtained under the same conditions as in FIG. 7). As shown in this chart, the thrust force F1 applied to the first bearing 31 is further reduced as compared with the first embodiment, and the force is not transmitted to the front wall of the case 10, so that the normal oil pump By comprising the case 10p, the load on the case front wall, which is disadvantageous in strength compared to other walls, can be eliminated. On the other hand, the thrust force F3 of the planetary gear set G2 is applied as it is to the third bearings 36 and 37 and the fourth bearing 38. However, as described above, these bearings are bearings having a small relative rotational difference. Will be small.

  Next, during the second speed (2nd) drive, the large-diameter first sun gear S2 also shares the reaction torque for power transmission for the reason described in the power transmission of the first embodiment. Therefore, the thrust force F2 acts. This thrust force is transmitted to the connecting member 12 via the bearing 34 and the second bearings 33 and 32, and as a result, an unbalanced force is generated on the speed reduction planetary gear G1 side that is balanced at the first speed. This force is applied as a load to the front wall of the case 10 via the first bearing 31 and the sun gear S1, but it is understood that this force is small as shown in FIG. In this case, the thrust force applied to the first bearing 31 is F1 + F2. On the other hand, on the planetary gear set G2 side, the thrust force F4 of the ring gear R3 is reduced by the torque share of the large-diameter sun gear S2 with respect to the thrust force F3 of the small-diameter sun gear S3. Is transmitted to the rear wall of the case via the fourth bearing 39. Of course, this force is the same as the force applied to the front wall of the case. In this case, the thrust force applied to the other three bearings 36, 37, and 38 is the same as that during the first speed drive.

  At the time of the next third speed (3rd) drive, the power transmission is different from the second speed only in that the large-diameter sun gear S2 is rotating, and in terms of torque sharing of elements involved in torque transmission. Since there is no particular difference, the relationship between the thrust forces is the second speed drive except that the value becomes smaller as the torque gain decreases with the reduction of the reduction ratio, as will be understood with reference to FIG. It will be exactly the same as time.

  Furthermore, power transmission during the fourth speed (4th) drive is performed in a state where torque transmission from the large-diameter sun gear S2 is lost with respect to the third speed. Accordingly, the thrust force relationship in this case is that the reduction planetary gear G1 and the planetary gear set G2 are both in the same closed loop as in the first speed drive, and no force is transmitted to the case 10, and the bearings 31, 36, 37, 38 The thrust force applied to is also reduced as the torque gain is reduced.

  Next, at the time of fifth speed (5th) drive shown in FIG. 14, the aspect of power transmission is different as described in the power transmission of the first embodiment, and the large-diameter sun gear S2 is connected to the output of the ring gear R3. On the other hand, the drive torque is received via the long pinion P2. As a result, the thrust force F2 applied to the large-diameter sun gear S2 is opposite to that at the time of the third speed drive, and this thrust force F2 is the bearing 35, the inner race of the one-way clutch, the carrier C2, It is transmitted to the ring gear R3 through both the bearings 37, 38 and the output connecting member 14, and balances with the thrust force F3 in the reverse direction of the ring gear R3. Therefore, the thrust force on the planetary gear set G2 side is a closed loop and does not exert an unbalanced force outside. On the other hand, on the deceleration planetary gear G1 side, the thrust force of the ring gear R1 and the sun gear S1 is reversed with respect to the first to fourth speed driving, so that the thrust force of the sun gear S1 is directly transmitted to the case front wall, The thrust force F1 of the ring gear R1 is transmitted to the case rear wall through the second bearings 32 and 33, the small-diameter sun gear S3, and the third and fourth bearings 36, 37, 38, and 39. Therefore, the thrust force F1 is applied to both the front and rear walls of the case 10 as a load. However, at this speed, the transmission torque is small due to the speed increase, and as is apparent with reference to FIG. 15, the thrust force F1 itself is extremely small compared to the second speed and the third speed, so Both the applied load and the bearing load are smaller than those at the gear positions.

  At the time of the next sixth speed (6th) drive, the power transmission is no longer on the deceleration planetary gear G1 side, so the thrust force due to the sun gear S1 and the ring gear R1 disappears compared to the fifth speed drive, and the planetary gear set G2 side Since the thrust force is balanced, the load on the case is eliminated.

  On the other hand, during reverse (Rev) driving, the power transmission is reverse rotation of the ring gear R3 output with respect to the input of the large-diameter sun gear S2, so that the thrust force F4 of the ring gear R3 is the thrust force F2 of the large-diameter sun gear S2. It becomes a relation away from. The thrust force F4 of the ring gear R3 in this case is transmitted from the output connecting member 14 to the case rear wall via the bearing 39. Further, the thrust force F2 of the large-diameter sun gear S2 is transmitted to the front wall of the case 10 via the sun gear S1 via the four bearings 34 to 31 on the front side. The thrust force F1 on the speed reduction planetary gear G1 side is balanced between the sun gear S1 and the ring gear R1. Therefore, during this backward drive, the thrust force F4 = F2 is transmitted to the front and rear walls of the case 10 together. In this case, both the bearing 31 and the bearing 39 receive a relatively large thrust force between the case 10 and the rotating members 12 and 14, but the bearing 39 is on the decelerating rotation side, so that it is not so large as a load. On the other hand, since the bearing 31 is on the input side, the relative rotational speed difference is large, so that the bearing load is larger than when the other gears are driven. Therefore, the bearing 31 needs to have a capacity suitable for the reverse drive, but the reverse drive itself occurs for a very short time during vehicle travel, so that it is extremely extreme to ensure the bearing durability immediately. It does not lead to the need for capacity.

  Next, FIG. 16 similarly shows, in a schematic cross section, a sixth embodiment in which the thrust bearing arrangement is changed with respect to the fifth embodiment. This form is intended to reduce the number of bearings. In the fifth embodiment, the bearing 35 disposed between the one-way clutch F-2 and the extended portion of the sun gear S2 is replaced with the two sun gears S2 of the planetary gear set G2. , S3 is used as a bearing 35 '. According to this configuration, the thrust force generated backward in the large-diameter sun gear S2 can be directly transmitted to the small-diameter sun gear S3 via the bearing 35 ', which is different from the conventional thrust force transmission via the carrier C2. Since the bearing 37 that serves to transmit the thrust force from the carrier C2 to the rear is not necessary, the number of bearings is reduced accordingly.

  In each of the above embodiments, the present invention has been embodied as a longitudinal transmission for an FR vehicle, and therefore a bearing arrangement that assumes an arrangement without a support wall that causes an increase in the axial length has been adopted. Is embodied in the form of a horizontal transmission for a front engine / front drive (FF) vehicle or a rear engine / rear drive (RR) vehicle, an output gear is provided in the normal transmission mechanism for parallel shaft output. A configuration in which the number of bearings is further reduced can be adopted in accordance with the need for a supporting support. Hereinafter, an embodiment of this type will be exemplified.

  17 to 23 show a seventh embodiment that takes the form of a three-axis transaxle. In FIG. 17, the overall configuration is shown as a skeleton with the shafts expanded, and in FIG. 18, the actual shaft arrangement relationship is shown in the axial direction. In this transmission, the main shaft X and the counter arranged in parallel to each other are shown. A three-axis configuration in which each element is arranged on each of the axis Y and the differential axis Z is employed. On the main shaft X, a torque converter 2 and a transmission mechanism having a gear train configuration substantially similar to that of the sixth embodiment are arranged, and on the counter shaft Y, a counter gear mechanism 4 also serving as a reduction mechanism is arranged. A differential device 5 is arranged on the differential shaft Z. In this embodiment, the parallel shaft output member on the main shaft X is a counter drive gear 19 'connected to the ring gear R3 as an output element of the planetary gear set G2. In addition, due to the restriction of the axial length associated with the horizontal installation, the engagement element that locks the large-diameter first sun gear S2 is only the band brake B-1. Therefore, although the names of the brake B-2 and the one-way clutch F-1 are raised, they correspond to the brake B-3 and the one-way clutch F-2 in the above-described embodiments.

  The counter gear mechanism 4 on the counter shaft Y is fixed to the counter shaft 40, and is fixed to the counter shaft 40 with a large-diameter counter driven gear 41 that meshes with a counter drive gear 19 ′ as an output member on the main shaft X. A small-diameter differential drive pinion gear 42 as an output element on the counter shaft Y is disposed, and by these, the output from the main shaft X side is decelerated on the parallel shaft, and transmitted to the differential device 5 by being reversed. In addition to obtaining an appropriate final reduction ratio, the rotation direction of the input shaft 11 and the output rotation direction from the differential device 5 are matched. The differential device 5 on the differential shaft Z is connected to the counter shaft 40 by meshing the differential ring gear 51 fixed to the differential case 52 with the differential drive pinion gear 42, and the differential gear disposed in the differential case 52 performs differential rotation. It is set as the structure output to the left-right axis | shaft 50, and this output is made into final wheel drive force.

  Also in this automatic transmission, as shown in the engagement chart of FIG. 19, the engagement and release of each clutch and brake (engaged with a circle, and released with a non-marked), and the speed stage achieved thereby. The relationship is the same as in each of the above embodiments (however, as described above, the brake B-2 and the one-way clutch F-1 correspond to the brake B-3 and the one-way clutch F-2 in each of the above embodiments). It becomes. In the figure, the circles in parentheses indicate the engagement during engine braking.

  FIG. 20 shows a schematic cross section of only the speed change mechanism on the main shaft X. In this speed change mechanism as well, the input shaft 11 rotates clockwise when the speed change mechanism is viewed from the torque converter side (the right side in the figure). Therefore, the pinion P1 of the planetary gear G1 for reduction is twisted counterclockwise, and the long pinion P2 of the planetary gear set G2 is twisted clockwise. This transmission mechanism is different from the previous embodiments in that a support wall 10 s fixed or integral with the case 10 is provided. The counter drive gear 19 'is supported on the support wall 10s via a radial ball bearing 18 capable of supporting thrust force.

  Each bearing in the transmission path includes a first bearing 31 between the sun gear S1 of the planetary gear G1 for reduction and the connecting member 12 that connects the ring gear R1 to the input shaft 11, and one second bearing 32 Between the connecting member 12 and the extension of the small diameter sun gear S3 of the planetary gear set G2, the other second bearing 34 is between the extension of the small diameter sun gear S3 and the extension of the large diameter sun gear S2, and the bearing. 35 'is between the two sun gears S2, S3, one fourth bearing 36 is between the small-diameter sun gear S3 and the other connecting member 13, and the other fourth bearing 38 is between the other connecting member 13 and the case. It is each arrange | positioned between the left walls.

  FIG. 21 and FIG. 22 show on a schematic cross section changes in thrust force at each gear position when the torsion of the above relationship is set. Referring to FIG. 21 (refer to FIG. 20 for the symbols indicating the respective members), at the time of the first speed (1st) drive, the speed reduction planetary gear G1 The thrust gear F1 acting on the right side in the figure acts on the ring gear R1, whereas the thrust force acting on the left side in the figure on the sun gear S1 acts on the ring gear R1. Therefore, the thrust force does not act on other bearings or the oil pump cover 10p as the case front wall. On the other hand, in the planetary gear set G2, a thrust force F3 to the left in the figure acts on the small-diameter sun gear S3, and an equivalent thrust force to the right in the figure acts on the ring gear R3. In this case, the thrust force of the small-diameter sun gear S3 is transmitted to the left wall of the case 10 via the bearing 36 and the bearing 38 of the transmission path, and the thrust force of the ring gear R3 is transmitted to the support 10s via the bearing 18. . Accordingly, when driving at the first speed (1st), the right wall, which is disadvantageous in strength due to the oil pump case 10p, is not loaded with a thrust load, and the thrust force generated only in the planetary gear set G2 is applied to the case 10. The left wall and the support wall 10s are loaded.

  FIG. 23 graphically shows the thrust force applied to each bearing during driving at each gear stage (in this table, numerical notation is omitted, but the values of the thrust forces F1 to F3 are shown in FIG. Equivalent to the value shown). As seen in this chart, the thrust force F1 applied to the bearing 31 is equivalent to that in the fifth embodiment. On the other hand, the thrust force F3 of the small-diameter sun gear S3 of the planetary gear set G2 is applied to the bearings 36 and 38 as they are. In this case, since the bearing 36 has a small relative rotational difference, the bearing 38 has a large relative rotational difference, so that the bearing load becomes large.

  Next, when driving at the second speed (2nd), the large-diameter sun gear S2 also shares the reaction torque for power transmission for the reason described in the power transmission of the first embodiment. Thrust force F2 acts. This thrust force is transmitted to the connecting member 12 via the second bearings 34 and 32. As a result, an unbalanced force is generated on the speed reduction planetary gear G1 side that is balanced at the first speed. The force F2 is applied to the case right wall as a load, but it can be seen that this force is small as shown in FIG. In this case, the thrust force applied to the bearing 31 is F1 + F2. On the other hand, on the planetary gear set G2 side, the thrust force F3 of the small-diameter sun gear S3 is transmitted to the case left wall via the bearings 36 and 38. Naturally, this force is the same as the sum of the thrust force F2 applied to the case right wall and the thrust force F4 applied to the support wall 10p. In this case, the thrust force applied to the other three bearings 36, 37, and 38 is the same as that during the first speed drive.

  At the time of the next third speed (3rd) drive, the power transmission is different from the second speed only in that the large-diameter sun gear S2 is rotating, and in terms of torque sharing of elements involved in torque transmission. Since there is no particular difference, the relationship between the thrust forces is the second speed drive except that the value becomes smaller as the torque gain decreases with the reduction of the reduction ratio, as will be understood with reference to FIG. It will be exactly the same as time.

  Furthermore, power transmission during the fourth speed (4th) drive is performed in a state where torque transmission from the large-diameter sun gear S2 is lost with respect to the third speed. Accordingly, the thrust force relationship in this case is that both the speed reduction planetary gear G1 and the planetary gear set G2 are in the same closed loop as in the first speed drive, and no force is transmitted to the case, and the bearings 31, 36, 37, 38 The applied thrust force is also reduced as the torque gain is reduced.

  Next, when driving at the fifth speed (5th) shown in FIG. 22, the aspect of power transmission is different as described in the power transmission of the first embodiment, and the large-diameter sun gear S2 is connected to the output of the ring gear R3. On the other hand, the drive torque is received via the long pinion P2. As a result, the thrust force F2 applied to the large-diameter sun gear S2 is opposite to that at the time of the third speed drive. Communicated. On the other hand, on the speed reduction planetary gear G1 side, the thrust force F1 of the ring gear R1 and the sun gear S1 is reversed with respect to the first to fourth speed driving, so that the thrust force of the sun gear S1 is directly transmitted to the case front wall. The thrust force of the ring gear R1 is transmitted to the case left wall via the bearing 32, the small-diameter sun gear S3, and the two bearings 36 and 38. Therefore, the thrust force F1 is applied to the right wall of the case, the thrust force F1 + F2 is applied to the left wall, and the thrust force F4 is applied to the support wall 10s. However, at this speed, the transmission torque is small due to the speed increase, and as is apparent with reference to FIG. 15, the thrust force itself is extremely small compared to the second speed and the third speed, so the load applied to the case Both bearing loads are smaller than those at those gear positions.

  At the time of the next sixth speed (6th) drive, the power transmission is no longer on the deceleration planetary gear G1 side, so the thrust force due to the sun gear S1 and the ring gear R1 disappears compared to the fifth speed drive, and the planetary gear set G2 side The thrust force is the same as in the fifth speed drive.

  On the other hand, during reverse (Rev) driving, the power transmission is reverse rotation of the ring gear R3 output with respect to the input of the large-diameter sun gear S2, so that the thrust force F4 of the ring gear R3 is the thrust force F2 of the large-diameter sun gear S2. It becomes a relation away from. The thrust force F4 of the ring gear R3 in this case is supported by the support wall 10s via the bearing 18. Further, the thrust force F2 of the large-diameter sun gear S2 is transmitted to the right wall of the case 10 via the sun gear S1 via the right three bearings 34, 32, 31. The thrust force F1 on the speed reduction planetary gear G1 side is balanced between the sun gear S1 and the ring gear R1. Therefore, during this reverse drive, no thrust load is applied to the left wall of the case 10, the thrust load F2 is applied to the right wall, and the thrust load F4 is supported by the support wall 10s. In this case, the bearing 31 receives a relatively large thrust force between the case 10 and the connecting member 12. Therefore, the bearing 31 needs to have a capacity suitable for the reverse drive, but the reverse drive itself occurs for a very short time during vehicle travel, so that it is extremely extreme to ensure the bearing durability immediately. It does not lead to the need for capacity.

  By the way, in the seventh embodiment, with regard to the configuration of the planetary gear set G2, the ring gear R3 is arranged to mesh with the long pinion P2 on the outer peripheral side of the large-diameter sun gear S2, but the ring gear R3 is arranged on the outer peripheral side of the small-diameter sun gear S3. It is also possible to adopt an arrangement that meshes with the long pinion P2. FIG. 24 shows a schematic cross section of a gear train according to an eighth embodiment having such an arrangement. Even if such a configuration is adopted, naturally, the same thrust force relationship as in the case of the seventh embodiment is established, and the same effect can be obtained.

  Finally, FIG. 25 shows the configuration of the planetary gear set G2, and the short pinion P3 and the ring gear R3 are increased in diameter compared to the seventh embodiment, the short pinion P3 is engaged with the small diameter sun gear S3 and the ring gear R3, and the long pinion P2 is increased. A ninth embodiment meshed with a sun gear S2 having a diameter is shown in a schematic cross section. Also in this case, since the generation of the rotational moment in the long pinion P2 can be prevented, the load on the rotation support portion can be reduced. Even if such a configuration is adopted, naturally, the same thrust force relationship as in the seventh and eighth embodiments is established, and the same effect can be obtained.

  The present invention has been described in detail with reference to nine embodiments. However, each of these embodiments is for illustrative purposes, and the present invention is within the scope of matters described in the individual claims. It can be implemented with various specific configurations being changed.

G1 Deceleration planetary gear G2 Planetary gear set S1 Sun gear (reaction force element)
P1 pinion C1 carrier (output element)
R1 Ring gear (one element of planetary gear for reduction)
S2 Sun gear (first sun gear)
S3 Sun gear (one element of planetary gear set, second sun gear)
P2 Long pinion P3 Short pinion C2, C3 Carrier R3, R2 Ring gear 10 Case 10p Oil pump case 11 Input shaft 12 Connecting member 13 Other connecting member 31 First bearing 32, 33 Second bearing 36, 37 Third bearing 38, 39 4th bearing

Claims (16)

Multistage shifting is achieved by a reduction planetary gear that is connected to the input shaft and outputs a reduction rotation to the output element by fixing the reaction force element, and a planetary gear set that outputs the reduction rotation by inputting the reduction rotation from the reduction planetary gear. In the automatic transmission to achieve,
Having at least a common transmission path through which a thrust force generated by one element of the planetary gear for deceleration and one element of the planetary gear set is transmitted at least when driving at the first speed;
The direction of the thrust force of the one element of the planetary gear for deceleration in the common transmission path and the direction of the thrust force of the one element of the planetary gear set are different from each other during the first speed drive. An automatic transmission characterized by setting a torsion direction of a helical tooth. The torsion direction is a direction in which the thrust force generated in the one element of the planetary gear for reduction and the one element of the planetary gear set faces each other during the first speed drive,
In the common transmission path, having a bearing that commonly receives the thrust force generated in each element,
The automatic transmission according to claim 1, wherein the bearing receives a differential thrust force between a thrust force acting on the one element of the planetary gear set and a thrust force acting on the one element of the planetary gear for reduction. The torsion direction is a direction in which the thrust force generated in the one element of the planetary gear for reduction and the one element of the planetary gear set is separated from each other during the first speed driving,
A bearing for receiving a thrust force generated in each element in the common transmission path;
The automatic transmission according to claim 1, wherein the bearing independently receives a thrust force acting on the one element of the planetary gear set and a thrust force acting on the one element of the speed reduction planetary gear. The reaction force element of the speed reduction planetary gear is fixed to the case,
The automatic transmission according to claim 2 or 3, wherein the thrust force transmitted to the common transmission path is transmitted to the case via the bearing.   The automatic transmission according to claim 2 or 3, wherein a carrier of the planetary gear set is supported in an axial direction on the common transmission path. The planetary gear set has a first sun gear that is one element thereof, and a second sun gear that is separate from the first sun gear,
The automatic transmission according to claim 2 or 3, wherein a bearing for receiving a thrust force acting on the second sun gear is disposed on a path different from the common transmission path. The planetary gear set has a first sun gear and a second sun gear separate from the first sun gear,
The automatic transmission according to claim 2 or 3, wherein a bearing is disposed between the first and second sun gears.   The automatic transmission according to claim 3, wherein the one element of the speed reduction planetary gear is a ring gear. The reaction force element of the speed reduction planetary gear is fixed to an oil pump case,
The automatic transmission according to claim 3 or 5, wherein a thrust force generated in the one element of the planetary gear for reduction is transmitted to the oil pump case via the bearing. The deceleration planetary gear is
A sun gear as the reaction force element;
A carrier connected to the one element of the planetary gear set supporting a pinion meshing with the sun gear;
A ring gear that meshes with a pinion supported by the carrier and is connected to the input shaft via a connecting member, which is the one element of the planetary gear for reduction.
The automatic transmission according to claim 1, wherein the bearing is disposed between the sun gear and the connecting member. The one element of the planetary gear set is a sun gear;
The connecting member is disposed between the speed reduction planetary gear and the planetary gear set,
The automatic transmission according to claim 10, wherein a second bearing is disposed between the connecting member and the sun gear.   The planetary gear set is engaged with one of a long pinion and a short pinion supported by a carrier, a first sun gear meshing with the long pinion, a second sun gear meshing with the short pinion, and a long pinion and a short pinion. The automatic transmission according to any one of claims 1 to 11, which is a Ravigneaux type planetary gear set comprising a ring gear.   The automatic transmission according to claim 12, wherein the ring gear meshes with the short pinion.   The automatic transmission according to claim 12, wherein the ring gear meshes with the long pinion.   The automatic transmission according to claim 14, wherein the long pinion meshes with the first sun gear at one end thereof, and the ring gear meshes with the other end of the long pinion. The input shaft is further connected to the planetary gear set carrier via another connecting member,
A third bearing is disposed between the sun gear of the planetary gear set and another connecting member;
A fourth bearing is disposed between the case and the other connecting member;
The automatic transmission according to claim 11, wherein the thrust force of the ring gear of the speed reduction planetary gear is transmitted to the case via the second, third, and fourth bearings during reverse drive.

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