JP2014181743A - Automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce friction loss by transmitting a thrust load generated by a gear of an automatic transmission to a transmission case so that the load goes through bearings as little as possible.SOLUTION: An automatic transmission includes thrust bearings T7 and T8 between a third sun gear Sc of a third planetary gear mechanism PGSc and a second sun gear Sb of a second planetary gear mechanism PGSb. Accordingly, a thrust load acting on the second sun gear Sb of the second planetary gear mechanism PGSb is transmitted from the thrust bearings T7 and T8 to an input shaft 12 via the third sun gear Sc of the third planetary gear mechanism PGSc, and the transmission of the thrust load from the third planetary gear mechanism PGSc to the input shaft 12 via a fourth planetary gear mechanism PGSd is thereby avoided as much as possible. As a result, friction loss can be reduced in a plurality of thrust bearings T9 to T12 lying between components of the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd, or the like.

Description

  In the present invention, a plurality of planetary gear mechanisms arranged on the axis of the input shaft are housed in a transmission case, and each element of the plurality of planetary gear mechanisms, the input shaft and the transmission case are connected. The present invention relates to an automatic transmission that establishes a predetermined gear stage by controlling with a plurality of clutches and a plurality of brakes.

  In such an automatic transmission, a device that absorbs differential rotation by disposing a plurality of thrust bearings between a plurality of rotating members that rotate relative to each other around the axis of an input shaft is known from Patent Document 1 below.

JP 2000-304107 A

  By the way, a helical gear with a small torque fluctuation is generally used as a gear of an automatic transmission, but the helical gear has a characteristic that an axial thrust load is generated by a meshing reaction force. Therefore, in an automatic transmission in which a plurality of thrust bearings are arranged between a plurality of rotating members that rotate relative to the axis of the input shaft, the friction of the thrust bearing that supports the thrust load increases, and the automatic transmission is configured in multiple stages. There is a possibility that the merit of widening and wide range may be lost.

  The present invention has been made in view of the above-described circumstances, and an object thereof is to reduce friction loss by transmitting a thrust load generated by a gear of an automatic transmission to a transmission case without passing through a bearing as much as possible. .

  In order to achieve the above object, according to the first aspect of the present invention, a plurality of planetary gear mechanisms arranged on the axis of the input shaft are housed in a transmission case, and the plurality of planetary gear mechanisms are arranged. In an automatic transmission that establishes a predetermined shift stage by controlling a coupling relationship between each element, the input shaft, and the transmission case with a plurality of clutches and a plurality of brakes, relative rotation with respect to the input shaft is impossible. A planetary gear mechanism A having a sun gear, a planetary gear mechanism B disposed adjacent to the planetary gear mechanism A and capable of rotating relative to each other, and the planetary gear mechanism B sandwiching the planetary gear mechanism A. A planetary gear mechanism C disposed on the opposite side and capable of rotating relative to each other; a clutch capable of coupling the ring gear of the planetary gear mechanism C to the input shaft; the sun gear of the planetary gear mechanism A and the planetary teeth A bearing A capable of supporting a thrust load disposed between sun gears of the mechanism B and a bearing B capable of supporting a thrust load disposed between a clutch hub of the clutch and the transmission case. An automatic transmission is proposed.

  According to the invention described in claim 2, in addition to the structure of claim 1, the first member extending radially inward from the ring gear of the planetary gear mechanism C, the clutch hub of the clutch, and the first A second member that is fixed to the input shaft so as not to rotate relative to the input shaft and extends radially outward; and a bearing C that can support a thrust load disposed between the first member and the second member. An automatic transmission characterized by this is proposed.

  According to the third aspect of the present invention, in addition to the configuration of the second aspect, the sun gear of the planetary gear mechanism B transmits the thrust load to the bearing A from the load transmission surface. An automatic transmission characterized in that a distance from C to a load receiving surface that receives a thrust load is larger than a total value of axial lengths of a plurality of components arranged between the load transmitting surface and the load receiving surface. Is proposed.

  According to the invention described in claim 4, in addition to the structure of any one of claims 1 to 3, the planetary gear mechanism C is fitted to the outer periphery of the input shaft so as to be relatively rotatable. A connecting member that connects the carrier of the planetary gear mechanism A to the carrier of the planetary gear mechanism A, and an outer peripheral spline that is fitted to an inner peripheral spline of the carrier of the planetary gear mechanism A is formed at one axial end of the connecting member. An automatic transmission is proposed in which an inner diameter of the inner peripheral spline is larger than an outer diameter of a sun gear of the planetary gear mechanism A.

  The second planetary gear mechanism PGSb of the embodiment corresponds to the planetary gear mechanism B of the present invention, and the third planetary gear mechanism PGSc of the embodiment corresponds to the planetary gear mechanism A of the present invention. The fourth planetary gear mechanism PGSd corresponds to the planetary gear mechanism C of the present invention, the third clutch C3 of the embodiment corresponds to the clutch of the present invention, and the seventh thrust bearing T7 and the eighth thrust bearing T8 of the embodiment. Corresponds to the bearing A capable of supporting the thrust load of the present invention, and the twelfth thrust bearing T12 of the embodiment corresponds to the bearing C capable of supporting the thrust load of the present invention, and the thirteenth thrust bearing T13 of the embodiment. Corresponds to the bearing B capable of supporting the thrust load according to the present invention, and the clutch drum 39 according to the embodiment corresponds to the first member according to the present invention. Strike plate 49 corresponds to the second member of the present invention.

  According to the configuration of the first aspect, the automatic transmission is disposed adjacent to the planetary gear mechanism A having the sun gear that cannot rotate relative to the input shaft, and the planetary gear mechanism A, and each element rotates relative to the planetary gear mechanism A. The planetary gear mechanism B that is possible, the planetary gear mechanism C that is disposed on the opposite side of the planetary gear mechanism B across the planetary gear mechanism A, and the ring gear of the planetary gear mechanism C are coupled to the input shaft. With possible clutches. Since the bearing A capable of supporting the thrust load disposed between the sun gear of the planetary gear mechanism A and the sun gear of the planetary gear mechanism B is provided, the thrust load acting on the sun gear of the planetary gear mechanism B is transferred from the bearing A to the planetary gear mechanism A. By transmitting to the input shaft via the sun gear, it is avoided as much as possible that the thrust load is transmitted from the planetary gear mechanism A to the input shaft via the planetary gear mechanism C, and the planetary gear mechanism A and the planetary gear mechanism C are avoided. It is possible to reduce the friction loss of a plurality of bearings interposed between these elements. Moreover, since the bearing B arranged between the clutch hub of the clutch and the transmission case capable of supporting the thrust load is provided, the thrust load transmitted to the input shaft can be supported by the transmission case via the bearing B. .

  According to the second aspect of the present invention, the first member extending radially inward from the ring gear of the planetary gear mechanism C and the clutch hub of the clutch and the first member are fixed to the input shaft so as not to rotate relative to each other. Since the second member extending outward and the bearing C capable of supporting the thrust load disposed between the first member and the second member are provided, the thrust load acting on the ring gear of the planetary gear mechanism C is directly applied to the input shaft. Can be transmitted and supported.

  According to the third aspect of the present invention, the distance from the load transmission surface where the sun gear of the planetary gear mechanism B transmits the thrust load to the bearing A to the load receiving surface where the second member receives the thrust load from the bearing C is the load transmission. A situation in which the thrust load of the sun gear of the planetary gear mechanism B is transmitted to the input shaft through the plurality of components because it is greater than the total axial length of the plurality of components disposed between the surface and the load receiving surface. Thus, friction can be reliably reduced.

  According to the fourth aspect of the present invention, the coupling member is fitted to the outer periphery of the input shaft so as to be relatively rotatable, and couples the carrier of the planetary gear mechanism C and the carrier of the planetary gear mechanism A, the axial direction of the coupling member One end is formed with an outer peripheral spline that fits into the inner peripheral spline of the carrier of the planetary gear mechanism A. Since the inner diameter of the inner peripheral spline is larger than the outer diameter of the sun gear of the planetary gear mechanism A, a connecting member is assembled in advance. When the planetary gear mechanism A is fitted to the outer periphery of the input shaft and the inner peripheral spline of the carrier of the planetary gear mechanism A is coupled to the outer peripheral spline of the connecting member, the inner peripheral spline of the carrier interferes with the sun gear of the planetary gear mechanism A. As a result, it is possible to avoid the situation where the assembly becomes impossible.

Skeleton diagram of automatic transmission. FIG. 2 is a detailed view of part 2 of FIG. 1. FIG. 3 is a detailed view of part 3 of FIG. The alignment chart of a 1st-4th planetary gear mechanism. Clutch and brake engagement table. Action | operation explanatory drawing at the time of the assembly of an automatic transmission. The table | surface which shows the direction of the torque of the sun gear of the 1st-4th planetary gear mechanism in each gear stage. The table | surface which shows the magnitude | size of the differential rotation of the 1st-13th thrust bearing in each gear stage. The table | surface which shows the magnitude | size of the thrust load of the 1st-13th thrust bearing in each gear stage. The table | surface which shows ratio of the thrust load of the 1st-4th planetary gear mechanism in each gear stage. The figure which shows the transmission path | route of thrust load F2, F3. The figure which shows the transmission path | route of thrust load F1, F2, F3. The figure which shows the transmission path | route of the thrust load F1, F2, F3, F4 in 6th speed, 7th speed, 9th speed, and 10th speed. The figure which shows the comparative example corresponding to FIG. The figure which shows the transmission path | route of the thrust load F1, F2, F3, F4 in the 3rd speed and 4th speed gear stage.

  Hereinafter, embodiments of the present invention will be described with reference to FIGS.

  First, the structure of the automatic transmission T having 10 forward speeds and 1 reverse speed will be described with reference to the skeleton diagram of FIG. 1 and the detailed views of FIG. 2 and FIG.

  The crankshaft 11 of the engine E is connected to the input shaft 12 of the automatic transmission T via a torque converter TC. On the outer periphery of the input shaft 12, the first planetary gear mechanism PGSa, the second planetary gear mechanism PGSb, the third planetary gear mechanism PGSc are arranged from the side closer to the engine E (right side in the figure) to the side farther (left side in the figure). The fourth planetary gear mechanism PGSd is arranged in order. In addition, in order to establish the respective gear positions by switching the coupling relationship of the elements of the first to fourth planetary gear mechanisms PGSa, PGSb, PGSc, and PGSd, the first clutch C1, the second clutch C2, the third clutch C3, 1 brake B1, 2nd brake B2, 3rd brake B3, and 4th brake B4 are provided. The second brake B2 is a two-way clutch, and the direction of relative rotation with which the second brake B2 is engaged can be arbitrarily switched. Other clutches and brakes are of a wet multi-plate type.

  The single-pinion type first planetary gear mechanism PGSa includes a first sun gear Sa, a first carrier Ca, a first ring gear Ra, and a plurality of first pinions Pa, and is rotatably supported by the first carrier Ca. The first pinions Pa ... mesh with the first sun gear Sa and the first ring gear Ra at the same time.

  The first sun gear Sa can be coupled to the transmission case 13 via the first brake B1. That is, the first brake B1 includes a brake hub 14 that is integral with the first sun gear Sa, a plurality of friction engagement elements 15 that are disposed between the brake hub 14 and the transmission case 13, and an axial slide on the transmission case 13. When the brake piston 16 is driven by hydraulic pressure supplied to the oil chamber 17 and the friction engagement elements 15 are engaged with each other, the first sun gear Sa is provided. Coupled to the transmission case 13.

  The transmission case 13 is actually composed of a plurality of members, but is illustrated as one member for convenience in the drawings. In the present specification, the stator shaft 19 fixed to the torque converter case 18 is also a part of the transmission case 13.

  The connecting member 20 extending radially outward from the first carrier Ca can be coupled to the transmission case 13 via the second brake B2. The first carrier Ca can be coupled to the input shaft 12 via the first clutch C1. That is, the first clutch C1 includes a clutch hub 22 connected to the radially inner end of the first carrier Ca through the connecting member 21, a clutch drum 23 fixed to the input shaft 12, and the clutch hub 22 and the clutch drum. Are provided between a plurality of friction engagement elements 24... And a clutch piston 25 slidably disposed inside the clutch drum 23. The clutch piston 25 is hydraulically supplied to the oil chamber 26. When the frictional engagement elements 24 are engaged with each other, the first carrier Ca is coupled to the input shaft 12.

  The first ring gear Ra is connected to a third carrier Cc of a third planetary gear mechanism PGSc, which will be described later, via a connecting member 27.

  The single pinion type second planetary gear mechanism PGSb includes a second sun gear Sb, a second carrier Cb, a second ring gear Rb, and a plurality of second pinions Pb... And is rotatably supported by the second carrier Cb. The second pinions Pb... Mesh with the second sun gear Sb and the second ring gear Rb at the same time.

  The second sun gear Sb is connected to a third ring gear Rc of a third planetary gear mechanism PGSc, which will be described later, via a connecting member 28. The second carrier Cb is integrally connected to the first carrier Ca of the first planetary gear mechanism PGSa via the connecting member 20, and can be coupled to the transmission case 13 via the second brake B2. An output gear 29 is formed integrally with the second ring gear Rb, and the output gear 29 is rotatably supported by the transmission case 13 via a pair of ball bearings 30 and 30.

  The single pinion type third planetary gear mechanism PGSc includes a third sun gear Sc, a third carrier Cc, a third ring gear Rc, and a plurality of third pinions Pc, and is rotatably supported by the third carrier Cc. The third pinions Pc... Mesh with the third sun gear Sc and the third ring gear Rc simultaneously.

  Third sun gear Sc is integrally coupled to input shaft 12. The third carrier Cc is connected to the first ring gear Ra of the first planetary gear mechanism PGSa via the connecting member 27. The third ring gear Rc can be coupled to a clutch drum 31 connected to a fourth sun gear Sd of a fourth planetary gear mechanism PGSd, which will be described later, via a second clutch C2. That is, the second clutch C2 includes a plurality of friction engagement elements 32 arranged between the third ring gear Rc and the clutch drum 31, and a clutch piston 33 slidably arranged inside the clutch drum 31. When the clutch piston 33 is driven by the hydraulic pressure supplied to the oil chamber 34 and the friction engagement elements 32 are engaged with each other, the third ring gear Rc of the third planetary gear mechanism PGSc becomes the fourth planetary gear mechanism. It is coupled to the fourth sun gear Sd of PGSd.

  The single pinion type fourth planetary gear mechanism PGSd includes a fourth sun gear Sd, a fourth carrier Cd, a fourth ring gear Rd, and a plurality of fourth pinions Pd, and is rotatably supported by the fourth carrier Cd. The fourth pinions Pd... Mesh with the fourth sun gear Sd and the fourth ring gear Rd simultaneously.

  The clutch drum 31 connected to the fourth sun gear Sd can be coupled to the transmission case 13 via the third brake B3. That is, the third brake B3 includes a plurality of friction engagement elements 35 arranged between the clutch drum 31 and the transmission case 13, and a brake piston 36 slidably arranged inside the transmission case 13. The fourth sun gear Sd is coupled to the transmission case 13 when the brake piston 36 is driven by the hydraulic pressure supplied to the oil chamber 37 to engage the friction engagement elements 35.

  The fourth carrier Cd is connected to the third carrier Cc of the third planetary gear mechanism PGSc via the connecting member 38. The fourth ring gear Rd can be coupled to the input shaft 12 via the third clutch C3. That is, the third clutch C3 includes a clutch drum 39 integrally connected to the fourth ring gear Rd, a clutch hub 40 integrally connected to the input shaft 12, and a plurality of clutches disposed between the clutch drum 39 and the clutch hub 40. Friction engagement elements 41... And a clutch piston 42 slidably disposed inside the clutch drum 39. The clutch piston 42 is driven by the hydraulic pressure supplied to the oil chamber 43 for friction engagement. When the elements 41 are engaged with each other, the fourth ring gear Rd is coupled to the input shaft 12.

  Further, the fourth ring gear Rd can be coupled to the transmission case 13 via the fourth brake B4. That is, the fourth brake B4 includes a plurality of friction engagement elements 44 disposed between the clutch drum 39 and the transmission case 13, and a brake piston 45 slidably disposed inside the transmission case 13. When the brake piston 45 is driven by the hydraulic pressure supplied to the oil chamber 46 and the friction engagement elements 44 are engaged with each other, the fourth ring gear Rd is coupled to the transmission case 13.

  The shaft end of the input shaft 12 on the engine E side is directly supported by a stator shaft constituting a part of the transmission case 13 via a ball bearing 47, and the shaft end of the input shaft 12 opposite to the engine E is located there. The fixed clutch hub 40 is supported by the transmission case 13 via a ball bearing 48.

  Each gear of the first planetary gear mechanism PGSa, the second planetary gear mechanism PGSb, the third planetary gear mechanism PGSc, and the fourth planetary gear mechanism PGSd is composed of a helical gear to reduce the variation in transmission torque, Since the axial thrust force acts by the force, 13 thrust bearings are arranged between the relatively rotating members.

  The first thrust bearing T1 is disposed between the stator shaft 19 constituting a part of the transmission case 13 and the clutch drum 23 of the first clutch C1. The second thrust bearing T2 is disposed between the clutch drum 23 and the clutch hub 22 of the first clutch C1. The third thrust bearing T3 is disposed between the clutch hub 22 and the first sun gear Sa of the first planetary gear mechanism PGSa. The fourth thrust bearing T4 is disposed between the first sun gear Sa and the first carrier Ca of the first planetary gear mechanism PGSa. The fifth thrust bearing T5 is disposed between the coupling member 27 connected to the first ring gear Ra of the first planetary gear mechanism PGSa and the second sun gear Sb of the second planetary gear mechanism PGSb. The sixth thrust bearing T6 is disposed between the connecting member 27 and the first carrier Ca of the first planetary gear mechanism PGSa.

  The seventh thrust bearing T7 is disposed between the second sun gear Sb of the second planetary gear mechanism PGSb and the third carrier Cc of the third planetary gear mechanism PGSc. The eighth thrust bearing T8 is disposed between the third carrier Cc and the third sun gear Sc of the third planetary gear mechanism PGSc. The ninth thrust bearing T9 is disposed between the third carrier Cc and the clutch drum 31 of the second clutch C2. The tenth thrust bearing T10 is disposed between the fourth sun gear Sd and the fourth carrier Cd of the fourth planetary gear mechanism PGSd. The eleventh thrust bearing T11 is disposed between the fourth carrier Cd and the clutch drum 39 of the third clutch C3. The twelfth thrust bearing T12 is disposed between the clutch drum 39 and a thrust plate 49 fixed to the input shaft 12. The thirteenth thrust bearing T13 is disposed between the clutch hub 40 of the third clutch C3 and the transmission case 13.

  A shim 50 (see FIG. 3) is disposed between the fourth sun gear Sd of the fourth planetary gear mechanism PGSd and the clutch drum 31 of the second clutch C2. The first carrier Ca of the first planetary gear mechanism PGSa and the connecting member 21 extending to the left in the drawing from the clutch hub 22 of the first clutch C1 are abutted by the abutting portion 51 (see FIG. 2). Between the first thrust bearing T1 and the clutch drum 23 of the first clutch C1, a shim 52 (see FIG. 2) for adjusting the dimensional error of each component stacked on the input shaft 12 is disposed. .

  The distance from the load transmitting surface P1 where the second sun gear Sb of the second planetary gear mechanism PGSb transmits the thrust load to the seventh thrust bearing T7 and the load receiving surface P2 where the thrust plate 49 receives the thrust load from the twelfth thrust bearing T12. L1 is a plurality of components arranged between the load transmission surface P1 and the load receiving surface P2, that is, the seventh thrust bearing T7, the third carrier Cc of the third planetary gear mechanism PGSc, the ninth thrust bearing T9, and the second clutch C2. Clutch drum 31, shim 50, fourth sun gear Sd of the fourth planetary gear mechanism PGSd, tenth thrust bearing T10, carrier Cd of the fourth planetary gear mechanism PGSd, eleventh thrust bearing T11, clutch drum 39 of the third clutch C3 And the axial length of the twelfth thrust bearing T12 It is set to be larger than the total value (see FIG. 3).

  An outer peripheral spline SPo formed on the outer peripheral surface of the right end in the drawing of the connecting member 38 connected to the fourth carrier Cd of the fourth planetary gear mechanism PGSd, and an inner periphery of the third carrier Cc of the third planetary gear mechanism PGSc in the left end of the drawing. The inner peripheral spline SPi formed on the surface meshes (see FIG. 6). The inner diameter r1 of the inner peripheral spline SPi of the third carrier Cc is set larger than the outer diameter r2 of the third sun gear Sc of the third planetary gear mechanism PGSc.

  The first carrier Ca of the first planetary gear mechanism PGSa receives a load from the load transmission surface P3 that transmits the thrust load to the fourth thrust bearing T4, and the clutch hub 22 of the first clutch C1 receives the thrust load from the third thrust bearing T3. The distance L2 to the surface P4 is set larger than the total value of the axial lengths of the fourth thrust bearing T4, the first sun gear Sa of the first planetary gear mechanism PGSa, and the third thrust bearing T3 (see FIG. 2). .

  FIG. 4 is a collinear diagram of the first to fourth planetary gear mechanisms PGSa, PGSb, PGSc, and PGSd. This corresponds to PGSa and the second planetary gear mechanism PGSb.

  For example, the gear ratio of the first planetary gear mechanism PGSa shown in the third stage from the top is h, the distance between the first sun gear Sa and the first carrier Ca, and the distance between the first carrier Ca and the first ring gear Ra. Is set to h: 1. Similarly, the gear ratio of the second planetary gear mechanism PGSb is i, the gear ratio of the third planetary gear mechanism PGSc is j, and the gear ratio of the fourth planetary gear mechanism PGSd is k.

  Of the two parallel horizontal lines in each collinear diagram, the upper horizontal line means that the rotational speed is “1” (the same rotational speed as the input shaft 12), and the lower horizontal line indicates that the rotational speed is “0”. ”(Stop). A speed line indicated by a broken line indicates that, among the first to fourth planetary gear mechanisms PGSa, PGSb, PGSc, and PGSd, other planetary gear mechanisms are idled following the planetary gear mechanism that transmits power.

  FIG. 7 shows the directions of torque acting on the first to fourth sun gears Sa, Sb, Sc, Sd of the first planetary gear mechanism PGSa, the second planetary gear mechanism PGSb, the third planetary gear mechanism PGSc, and the fourth planetary gear mechanism PGSd. “+” Indicates that the torque direction of the engine E is the same, and “−” indicates that the torque direction of the engine E is opposite. The direction of the thrust load acting on the first to fourth sun gears Sa, Sb, Sc, Sd made of helical gears is determined according to the direction of torque.

  On the first sun gear Sa of the first planetary gear mechanism PGSa, a leftward thrust load acts in the first to fifth gears, and no thrust load acts on the other gears. On the second sun gear Sb of the second planetary gear mechanism PGSb, a leftward thrust load in the drawing acts on all the forward gears, and a rightward thrust load on the reverse gears acts. On the third sun gear Sc of the third planetary gear mechanism PGSc, a leftward thrust load acts in the forward shift speed except the 8-speed shift speed, and a rightward thrust load acts in the reverse speed speed. The fourth sun gear Sd of the fourth planetary gear mechanism PGSd is subjected to a leftward thrust load in the figure at the sixth speed, the seventh speed, the ninth speed and the tenth speed, and the third speed, A rightward thrust load acts in the speed shift stage and the reverse shift stage.

  FIG. 8 shows an example of differential rotation of the first to thirteenth thrust bearings T <b> 1 to T <b> 13 at each gear stage, and 1,000 rpm corresponds to differential rotation between the input shaft 12 and the transmission case 13. For example, the differential rotation between the first thrust bearing T1 and the thirteenth thrust bearing T13 disposed between the input shaft 12 and the transmission case 13 is 1,000 rpm.

  FIG. 9 shows an example of the thrust load acting on the first to thirteenth thrust bearings T1 to T13 at each shift stage. Since the thrust load acting on the first to thirteenth thrust bearings T1 to T13 varies depending on the torsion angle of the gear, an example thereof is shown in FIG.

  Since the sun gear and the ring gear of the single pinion type planetary gear mechanism mesh with the common pinion, the thrust loads received by the sun gear and the ring gear are the same and opposite to each other. FIG. 10 shows the thrust load F1 of the first sun gear Sa and the first ring gear Ra of the first planetary gear mechanism PGSa, the thrust load F2 of the second sun gear Sb and the second ring gear Rb of the second planetary gear mechanism PGSb, and the third planetary gear. The thrust load F3 of the third sun gear Sc and the third ring gear Rc of the gear mechanism PGSc, and the thrust load F4 of the fourth sun gear Sd and the fourth ring gear Rd of the fourth planetary gear mechanism PGSd are shown for each shift stage. . In addition, the numerical value of F1-F4 in FIG. 10 has shown the relative magnitude | size between them.

  Next, an embodiment of the present invention having the above configuration will be described.

  First, the torque flow of each gear stage will be described based on the skeleton diagram of FIG. 1, the alignment chart of FIG. 4, and the engagement table of FIG. In the engagement table of FIG. 5, the circles indicate that the clutch or brake is in an engaged state. Further, since the second brake B2 is constituted by a two-way clutch, the second brake B2 is switched between the forward rotation prevention state F and the reverse rotation prevention state R. Underlined F and R indicate that the rotational speeds of the first carrier Ca and the second carrier Cb become “0” by the action of the second brake B2.

  When the first gear is established, the first brake B1 is engaged, the second brake B2 is set in the reverse rotation prevention state, and the third brake B3 is engaged. By setting the second brake B2 to the reverse rotation prevention state, the reverse rotation of the first carrier Ca and the second carrier Cb is prevented. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. As a result, the three elements of the first planetary gear mechanism PGSa enter a locked state in which relative rotation is impossible, and the rotational speeds of the first ring gear Ra, the third carrier Cc, and the fourth carrier Cd also become “0”. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “1st” shown in FIG. 4, and the first gear is established.

  It is not necessary to engage the third brake B3 in order to establish the first gear, but if the third brake B3 is engaged in advance while the first gear is established, the first gear is changed. The transition to the second gear can be smoothly performed. If it is necessary to operate the engine brake at the first gear, the second brake B2 may be switched from the reverse rotation prevention state to the forward rotation prevention state.

  When the second gear is established, the first brake B1 is engaged, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, and the second clutch C2 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the second clutch C2, the rotation speeds of the third ring gear Rc and the second sun gear Sb become “0”, which is the same as the rotation speed of the fourth sun gear Sd. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “2nd” shown in FIG. 4, and the second gear is established.

  When the third speed is established, the first brake B1 is engaged, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, and the third clutch C3 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the third clutch C <b> 3, the rotational speed of the fourth ring gear Rd becomes “1”, which is the same speed as the rotational speed of the third sun gear Sc connected to the input shaft 12. In this way, the rotational speed of the fourth sun gear Sd becomes “0” and the rotational speed of the fourth ring gear Rd becomes “1”, so the rotational speed of the fourth carrier Cd, that is, the fourth carrier Cd, the third The rotational speeds of the carrier Cc and the first ring gear Ra are k / (k + 1). As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “3rd” shown in FIG. 4, and the third gear is established.

  When the 4th speed is established, the first brake B1 is engaged, the second brake B2 is set in the reverse rotation prevention state, the second clutch C2 is engaged, and the third clutch C3 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. By engaging the second clutch C2, the fourth sun gear Sd, the third ring gear Rc, and the second sun gear Sb rotate at the same speed. As a result, the third carrier Cc and the fourth carrier Cd are connected between the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd, and the third ring gear Rc and the fourth sun gear Sd are connected. In the fourth speed gear stage that engages the second clutch C2, the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd can draw one collinear diagram composed of four elements.

  By engaging the third clutch C3, the rotation speed of the fourth ring gear Rd becomes “1”, which is the same as the rotation speed of the third sun gear Sc, and the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd. The rotational speed of two of the four elements composed of “1” becomes “1” of the same speed. As a result, the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd are in a locked state where the respective elements cannot rotate relative to each other, and the rotational speeds of all the elements of the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd are It becomes “1”. Then, the rotational speeds of the first carrier Ca and the second carrier Cb become h / (h + 1), and the rotational speed of the second ring gear Rb to which the output gear 29 is connected becomes “4th” shown in FIG. A stage is established.

  When the fifth gear is established, the first brake B1 is engaged, the second brake B2 is in the reverse rotation prevention state, the first clutch C1 is engaged, and the third clutch C3 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “5th” shown in FIG. 4, and the fifth gear is established.

  It is not necessary to engage the third clutch C3 when establishing the fifth gear, but the third clutch C3 needs to be engaged at the fourth and sixth gears adjacent to the fifth gear. Therefore, by engaging the third clutch C3 while establishing the fifth gear, the shift change to the fourth gear or the sixth gear can be performed smoothly.

  When the sixth speed is established, the second brake B2 is set in the reverse rotation prevention state, the first clutch C1, the second clutch C2, and the third clutch C3 are engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the second clutch C2 and the third clutch C3, the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd are brought into the locked state as described in the fourth speed, and the third ring gear Rc. The rotational speed of the second sun gear Sb is “1”. By engaging the first clutch C1, the rotation speeds of the first carrier Ca and the second carrier Cb become “1”.

  Thereby, in the second planetary gear mechanism PGSb, the rotation speeds of the second carrier Cb and the second sun gear Sb become “1” of the same speed, and each element is in a locked state in which the relative rotation is impossible. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “1” of “6th” shown in FIG. 4, and the sixth speed gear stage is established.

  When the seventh speed is established, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, the first clutch C1 is engaged, and the third clutch C3 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the third clutch C3, the rotational speed of the fourth ring gear Rd becomes “1”, which is the same as the rotational speed of the third sun gear Sc, and the third carrier Cc, the first ring gear Ra, and the fourth carrier Cd. Becomes the rotation speed k / (k + 1). By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc connected to the input shaft 12. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “7th” shown in FIG. 4, and the seventh speed gear stage is established.

  At the time of establishment of the eighth gear, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, the first clutch C1 is engaged, and the second clutch C2 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the second clutch C2, the rotational speeds of the third ring gear Rc and the second sun gear Sb become “0”, which is the same speed as the rotational speed of the fourth sun gear Sd. By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “8th” shown in FIG. 4, and the 8-speed gear stage is established.

  When the ninth gear is established, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, the fourth brake B4 is engaged, and the first clutch C1 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the fourth brake B4, the rotational speed of the fourth ring gear Rd is also "0". As a result, the fourth planetary gear mechanism PGSd is locked, and the rotational speeds of the fourth carrier Cd, the third carrier Cc, and the first ring gear Ra are also “0”.

  By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “9th” shown in FIG. 4, and the ninth gear is established.

  When the tenth speed is established, the second brake B2 is set in the reverse rotation prevention state, the fourth brake B4 is engaged, the first clutch C1 is engaged, and the second clutch C2 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the second clutch C2, the third ring gear Rc and the second sun gear Sb rotate at the same speed as the fourth sun gear Sd. By engaging the fourth brake B4, the rotational speed of the fourth ring gear Rd becomes “0”. By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “10th” shown in FIG.

  At the time of establishment of the reverse gear, the second brake B2 is set in the forward rotation blocking state, the third brake B3 is engaged, and the third clutch C3 is engaged. By engaging the third brake B3 and the third clutch C3, the rotational speeds of the third carrier Cc, the first ring gear Ra, and the fourth carrier Cd become k / (k + 1). By setting the second brake B2 in the forward rotation blocking state, the forward rotation of the first carrier Ca and the second carrier Cb is blocked, and the rotation speeds of the first carrier Ca and the second carrier Cb become “0”. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “Rvs” shown in FIG. 4, and the reverse gear position is established.

  Next, reduction of thrust bearing friction will be described.

  The second sun gear Sb of the second planetary gear mechanism PGSb is subjected to a leftward thrust load F2 in the drawing at all forward speeds (see FIGS. 7 and 11). The thrust load F2 is transmitted to the input shaft 12 via the seventh thrust bearing T7, the carrier Cc of the third planetary gear mechanism PGSc, the eighth thrust bearing T8, and the third sun gear Sc of the third planetary gear mechanism PGSc, and the input shaft 12 To the transmission case 13 via the clutch hub 40 of the third clutch C3 and the thirteenth thrust bearing T13.

  At that time, as shown in FIG. 3, the second sun gear Sb of the second planetary gear mechanism PGSb transmits the thrust load to the seventh thrust bearing T7, and the thrust plate 49 extends from the twelfth thrust bearing T12. The distance L1 to the load receiving surface P2 that receives the thrust load is the seventh thrust bearing T7 disposed between the load transmitting surface P1 and the load receiving surface P2, the third carrier Cc of the third planetary gear mechanism PGSc, and the ninth thrust bearing. T9, clutch drum 31 of the second clutch C2, shim 50, fourth sun gear Sd of the fourth planetary gear mechanism PGSd, tenth thrust bearing T10, carrier Cd of the fourth planetary gear mechanism PGSd, eleventh thrust bearing T11, third Clutch drum 39 of clutch C3 and twelfth thrust bearing T12 Since the axial length is set to be larger than the total value, the thrust load F2 of the second sun gear Sb of the second planetary gear mechanism PGSb is transmitted to the transmission case 13 via the input shaft 12 through the shortest path. become. As a result, the thrust load F2 is prevented from passing through many parts such as the fourth planetary gear mechanism PGSd and the thrust bearings T9 to T12, and the friction can be reduced and the parts can be made thinner.

  Further, at the third speed gear stage and the fourth speed gear stage, a leftward thrust load F4 (see FIGS. 7 and 15) acts on the fourth ring gear Rd of the fourth planetary gear mechanism PGSd. The transmission can be transmitted from the 4-ring gear Rd to the input shaft 12 via the clutch drum 39, the twelfth thrust bearing T12 and the thrust plate 49, and supported by the transmission case 13 via the thirteenth thrust bearing T13.

  FIG. 6 shows a procedure for determining the thickness of the shim 50 (see FIG. 3). First, the third clutch C3, the fourth brake B4, the third brake B3, the fourth planetary gear mechanism PGSd, and the like are assembled on the outer periphery of the input shaft 12. In this state, from the surface of the fourth planetary gear mechanism PGSd that contacts the shim 50 of the fourth sun gear Sd to the right end of the third sun gear Sc of the third planetary gear mechanism PGSc that is pre-splined to the input shaft 12 in the drawing. Measure the distance X.

  Next, the distance Y from the left end (surface contacting the shim 50) of the clutch drum 31 of the second clutch C2 assembled in advance as a subassembly to the right end (surface contacting the ninth thrust bearing T9) in the drawing is measured. To do. The third carrier Cc, the third pinion Pc, the eighth thrust bearing T8, and the ninth thrust bearing T9 are preassembled as subassemblies, and the left end of the ninth thrust bearing T9 in the figure and the left end of the eighth thrust bearing T8 in the figure. The distance Z to is measured. Then, a value slightly smaller than X− (Y + Z) is set as the thickness of the shim 50, and the shim 50 corresponding to the above-described thickness may be selected from a plurality of types of shim 50 prepared in advance. .

  As described above, when the pre-assembled sub-assembly of the third carrier Cc is assembled so as to be fitted to the outer periphery of the input shaft 12, the inner peripheral spline SPi of the third carrier Cc extends the outer periphery of the third sun gear Sc. Although passing from the right side to the left side in the drawing, the radius r1 of the inner peripheral spline SPi of the third carrier Cc is set to be larger than the radius r2 of the third sun gear Sc, so that the third carrier Cc does not interfere with the third sun gear Sc. A subassembly of the planetary gear mechanism PGSc can be assembled.

  As described above, the gap between the ring gear Rc of the third planetary gear mechanism PGSc and the sun gear Sd of the fourth planetary gear mechanism PGSd can be coupled to the clutch drum 31 of the third clutch C3 and the sun gear Sd of the fourth planetary gear mechanism PGSd. Since the shim 50 is disposed in the position, the insertion of the shim 50 does not affect the relative distance between the elements of the fourth planetary gear mechanism PGSd or the distance between the elements of the second clutch C2. Since the shim 50 can be arranged using the dead space between the planetary gear mechanism PGSd and the second clutch C2, an increase in the axial dimension of the automatic transmission T can be minimized.

  As shown in FIG. 2, the left-side end of the first carrier Ca of the first planetary gear mechanism PGSa and the left-side end of the connecting member 21 extending from the clutch hub 22 of the first clutch C1 are shown. The thrust load is applied to the second sun gear Sb of the second planetary gear mechanism PGSb when the vehicle travels at a reduced speed (when the engine brake is activated) or when the vehicle travels backward. The fifth thrust bearing T5 → the connecting member 27 → the sixth thrust bearing T6 → the carrier Ca of the first planetary gear mechanism PGSa → the abutting portion 51 → the connecting member 21 → the clutch hub 22 of the first clutch C1 → the second thrust bearing. Stator shaft 19 (transmission case 1) in the path of T2 → clutch drum 23 of first clutch C1, shim 52 → first thrust bearing T1. ) Is transmitted to the.

  At this time, if the abutting portion 51 does not exist, the load input from the sixth thrust bearing T6 to the carrier Ca of the first planetary gear mechanism PGSa does not pass through the connecting member 21, and the fourth thrust bearing T4 → The sun gear Sa of the first planetary gear mechanism PGSa is transmitted from the sun gear Sa of the first planetary gear mechanism PGSa to the clutch hub 22 of the first clutch C1 through the path of the third thrust bearing T3 and then to the second thrust bearing T2, and is rotated at a high speed. Large friction is generated in the fourth thrust bearing T4 and the third thrust bearing T3 on both sides of the first and second thrust bearings. Moreover, when the thrust load passes through the fourth thrust bearing T4 and the third thrust bearing T3, the thickness type of the shim 52 increases in order to compensate for the dimensional tolerance of the fourth thrust bearing T4 and the third thrust bearing T3. There is a problem.

  On the other hand, according to the present embodiment, since the abutment portion 51 is provided, the thrust load bypasses the fourth thrust bearing T4 and the third thrust bearing T3 that sandwich the first sun gear Sa of the first planetary gear mechanism PGSa. In addition to reducing the friction of the fourth thrust bearing T4 and the third thrust bearing T3, it is not necessary to consider these dimensional tolerances, and the thickness type of the shim 52 can be reduced.

  At this time, the first carrier Ca of the first planetary gear mechanism PGSa transmits a thrust load from the third thrust bearing T3 to the clutch hub 22 of the first clutch C1 from the load transmission surface P3 that transmits the thrust load to the fourth thrust bearing T4. The distance L2 to the load receiving surface P4 to be received is larger than the total value of the axial lengths of the fourth thrust bearing T4, the first sun gear Sa of the first planetary gear mechanism PGSa, and the third thrust bearing T3. To avoid the situation where the rightward thrust load input from the gear mechanism PGSb is transmitted to the transmission case 13 via the fourth thrust bearing T4 and the third thrust bearing T3, the friction can be reduced reliably. it can.

  The end portion of the first carrier Ca of the first planetary gear mechanism PGSa provided with the abutting portion 51 also serves as a holding portion for the sixth thrust bearing T6.

  Next, the setting of the direction of the helical angle of the helical gears of the first planetary gear mechanism PGSa, the second planetary gear mechanism PGSb, the third planetary gear mechanism PGSc, and the fourth planetary gear mechanism PGSd will be described.

  Since the second ring gear Rb of the second planetary gear mechanism PGSb is formed integrally with the output gear 29, the rightward thrust load F2 generated in the figure by the second ring gear Rb is transmitted via the output gear 29 and the ball bearings 30 and 30. And transmitted to the transmission case 13. Therefore, the leftward thrust load F2 generated by the second sun gear Sb of the second planetary gear mechanism PGSb directly cancels the rightward thrust load F2 generated by the second ring gear Rb of the second planetary gear mechanism PGSb. However, it is necessary to transmit and support the transmission case 13 by any route. At this time, if the thrust load F2 facing left in the drawing of the second sun gear Sb can be offset with the thrust load generated by another planetary gear mechanism, the thrust bearing existing in the path transmitted to the transmission case 13 will be described. Friction can be reduced.

  From this point of view, in the present embodiment, as shown in FIG. 11, the leftward thrust load F2 generated by the second sun gear Sb of the second planetary gear mechanism PGSb is applied to the third planetary gear mechanism PGSc adjacent thereto. The rightward thrust load F3 generated by the third ring gear Rc and the connecting member 28 immediately cancel each other, and only the thrust load F2-F3, which is the difference between the thrust load F3 and the third planetary gear mechanism PGSc, is offset. The transmission is transmitted to the transmission case 13 via the 3 sun gear Sc, the input shaft 12 and the thirteenth thrust bearing T13.

  As is apparent from FIG. 10, in this embodiment, the thrust load F2 facing left in the drawing of the second sun gear Sb is larger than the thrust load F3 facing right in the drawing of the third ring gear Rc in all forward speeds. The thrust load F2-F3 that is the difference is always leftward in the figure. However, the above-described relationship of F2> F3 does not always hold, and the gear twist angle may be set so that F2 <F3 at a low speed gear (for example, the first gear or the second gear). . This is because, at low speed gears used for starting and running at low speed, the loss of power transmission is originally large, so that the effect of reducing friction obtained is small even if the relationship of F2> F3 is established.

  A leftward thrust load F2 of the second sun gear Sb in the figure is input to the seventh thrust bearing T7 and the eighth thrust bearing T8, and a part of the thrust load F2 is a rightward thrust load of the third ring gear Rc in the figure. By canceling with F3, the load on the seventh thrust bearing T7 and the eighth thrust bearing T8 can be reduced.

  In FIG. 11, it is assumed that the first planetary gear mechanism PGSa and the fourth planetary gear mechanism PGSd are unloaded.

  As shown in FIGS. 7 and 12, the thrust load F1 acts on the first sun gear Sa and the first ring gear Ra of the first planetary gear mechanism PGSa only in the first gear to the fifth gear, and the first The leftward thrust load F1 of the sun gear Sa in the drawing is equal to the rightward thrust load F1 of the first ring gear Ra in the drawing. Then, the leftward thrust load F1 of the first sun gear Sa in the drawing is transmitted to the first ring gear Ra through the path of the fourth thrust bearing T4 → the first carrier Ca → the sixth thrust bearing T6 → the connecting member 27, and the first ring gear Ra. The thrust load F1 is completely offset with the rightward thrust load F1, so that the thrust load F1 does not affect other planetary gear mechanisms.

  The direction of the thrust load acting on the sun gear of the planetary gear mechanism is determined by the rotation direction (torque direction) of the sun gear and the direction of the torsion angle of the tooth surface, but the direction of the torsion angle does not change. The direction of thrust load changes according to the change of. In the single pinion type planetary gear mechanism, the thrust load acting on the sun gear and the thrust load acting on the ring gear are directed in opposite directions.

  In the present embodiment, since the rotation direction (torque direction) of the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is as shown in FIG. 7, the fourth sun gear Sd and the fourth ring gear of the fourth planetary gear mechanism PGSd. The thrust load F4 of Rd is rightward in the drawing and leftward in the drawing at the third gear, the fourth gear, and the reverse gear, respectively, but the sixth gear, the seventh gear, the ninth gear, and the tenth gear. At the speed shift stage, they change to the left in the figure and to the right in the figure, respectively. By setting the direction of the thrust load F4 of the fourth sun gear Sd and the fourth ring gear Rd as described above, the following operational effects can be obtained.

  As shown in FIG. 8, in the 9th speed stage and the 10th speed stage, which are the high speed speed stages, there is a problem that the differential rotation of the 12th thrust bearing T12 reaches 1000 rpm at the maximum and the friction increases. By setting the direction of the thrust load F4 of Sd and the fourth ring gear Rd as described above, the fourth sun gear Sd and the fourth ring gear Rd of the ninth speed gear and the tenth speed gear as shown in FIG. The thrust load F4 is canceled by the tenth thrust bearing T10 and the eleventh thrust bearing T11, and is not transmitted to the twelfth thrust bearing T12 having a large differential rotation, and the friction of the twelfth thrust bearing T12 can be effectively reduced.

  By the way, at the sixth speed, the rightward thrust load F4 acting on the fourth sun gear Sd of the fourth planetary gear mechanism PGSd increases (see FIG. 10). Accordingly, assuming that the direction of the thrust load F4 acting on the fourth ring gear Rd and the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is set opposite to the embodiment described in FIG. 13, as shown in FIG. Furthermore, the left thrust load F2-F3 in the drawing and the right thrust load F4 in the drawing transmitted from the fourth sun gear Sd through the ninth thrust bearing T9 are input to the seventh thrust bearing T7. At the sixth speed, the rightward thrust load F4 in the figure increases, so the resultant force F2-F3-F4 has a negative value, and the seventh thrust bearing T7 may be urged to the right in the figure. .

  In such a situation, the thrust load for urging the seventh thrust bearing T7 in the right direction in the figure is the seventh thrust bearing T7 → the fifth thrust bearing T5 → the sixth thrust bearing T6 → the second thrust bearing T2 → the first thrust. This is transmitted to the stator shaft 19 (transmission case 13) through the path of the bearing T1, and the friction of the five thrust bearings T7 to T5, T2, and T1 existing in the path increases.

  From the above, at the sixth speed, the seventh speed, the ninth speed and the tenth speed, the thrust load F4 of the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is directed leftward in the figure. It is necessary to set the helix angle of the helical gear.

  By the way, as described above, the thrust load F4 of the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is directed leftward in the drawing at the sixth speed, the seventh speed, the ninth speed, and the tenth speed. When the helical gear torsion angle is set, the direction of the thrust load F4 of the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is reversed at the third speed gear stage and the fourth speed gear stage and becomes rightward in the figure (see FIG. 7). As described in FIG. 14, the thrust load F4 of the fourth sun gear Sd may be transmitted to the five thrust bearings T7 to T5, T2, and T1.

  However, the rightward thrust load F4 of the fourth sun gear Sd in the drawing at the third and fourth gears is smaller than the rightward thrust load F4 of the fourth sun gear Sd in the sixth gear (see the figure). 10), the resultant force F2-F3-F4 does not become negative, and the thrust load F4 is transmitted to the stator shaft 19 (transmission case 13) via five thrust bearings T7 to T5, T2, T1. Can be avoided (see FIG. 15).

  Further, in the third gear and the fourth gear, the leftward thrust load F4 of the fourth ring gear Rd of the fourth planetary gear mechanism PGSd is transmitted to the twelfth thrust bearing T12, but the third gear and the fourth gear. Since the differential rotation of the twelfth thrust bearing T12 is zero at the shift speed (see FIG. 8), there is no problem that the friction of the twelfth thrust bearing T12 increases even if the thrust load F4 is input.

  The embodiments of the present invention have been described above, but various design changes can be made without departing from the scope of the present invention.

  For example, although the third sun gear Sc of the third planetary gear mechanism PGSc is splined to the input shaft 12 in the embodiment, the third sun gear Sc may be formed integrally with the input shaft 12.

  In the embodiment, the second brake B2 is constituted by a two-way clutch, but it can be constituted by a multi-number type brake.

  In the embodiment, the thirteenth thrust bearing T13 transmits the thrust load input from the input shaft 12 to the transmission case 13 through the outer ring of the ball bearing 48. However, the thirteenth thrust bearing T13 is abolished and the ball The thrust load input from the input shaft 12 may be transmitted to the transmission case 13 by changing the bearing 48 to a ball bearing (for example, an angular ball bearing) that can also support a thrust load.

PSGb Second planetary gear mechanism (planetary gear mechanism B)
Sb Sun gear PSGc of planetary gear mechanism B Third planetary gear mechanism (planetary gear mechanism A)
Sc Sun gear Cc of planetary gear mechanism A Carrier PSGd of planetary gear mechanism A Fourth planetary gear mechanism (planetary gear mechanism C)
Cd Carrier Rd of planetary gear mechanism C Ring gear C3 of planetary gear mechanism C Third clutch (clutch)
T7 7th thrust bearing (bearing A capable of supporting thrust load)
T8 8th thrust bearing (bearing A capable of supporting thrust load)
T12 12th thrust bearing (bearing C capable of supporting thrust load)
T13 13th thrust bearing (bearing B capable of supporting thrust load)
SPi Planetary gear mechanism A carrier inner peripheral spline SPo Connecting member outer peripheral spline r1 Inner peripheral spline inner diameter r2 Planetary gear mechanism A sun gear outer diameter P1 Load transmission surface P2 Load receiving surface L1 From load transmitting surface to load receiving surface Distance 12 Input shaft 13 Transmission case 38 Connecting member 39 Clutch drum (first member)
40 Clutch hub 49 Thrust plate (second member)

  In the present invention, a plurality of planetary gear mechanisms arranged on the axis of the input shaft are housed in a transmission case, and each element of the plurality of planetary gear mechanisms, the input shaft and the transmission case are connected. The present invention relates to an automatic transmission that establishes a predetermined gear stage by controlling with a plurality of clutches and a plurality of brakes.

  In such an automatic transmission, a device that absorbs differential rotation by disposing a plurality of thrust bearings between a plurality of rotating members that rotate relative to each other around the axis of an input shaft is known from Patent Document 1 below.

JP 2000-304107 A

  By the way, a helical gear with a small torque fluctuation is generally used as a gear of an automatic transmission, but the helical gear has a characteristic that an axial thrust load is generated by a meshing reaction force. Therefore, in an automatic transmission in which a plurality of thrust bearings are arranged between a plurality of rotating members that rotate relative to the axis of the input shaft, the friction of the thrust bearing that supports the thrust load increases, and the automatic transmission is configured in multiple stages. There is a possibility that the merit of widening and wide range may be lost.

  The present invention has been made in view of the above-described circumstances, and an object thereof is to reduce friction loss by transmitting a thrust load generated by a gear of an automatic transmission to a transmission case without passing through a bearing as much as possible. .

  In order to achieve the above object, according to the first aspect of the present invention, a plurality of planetary gear mechanisms arranged on the axis of the input shaft are housed in a transmission case, and the plurality of planetary gear mechanisms are arranged. In an automatic transmission that establishes a predetermined shift stage by controlling a coupling relationship between each element, the input shaft, and the transmission case with a plurality of clutches and a plurality of brakes, relative rotation with respect to the input shaft is impossible. A planetary gear mechanism A having a sun gear, a planetary gear mechanism B disposed adjacent to the planetary gear mechanism A and capable of rotating relative to each other, and the planetary gear mechanism B sandwiching the planetary gear mechanism A. A planetary gear mechanism C disposed on the opposite side and capable of rotating relative to each other; a clutch capable of coupling the ring gear of the planetary gear mechanism C to the input shaft; the sun gear of the planetary gear mechanism A and the planetary teeth A bearing A capable of supporting a thrust load disposed between sun gears of the mechanism B and a bearing B capable of supporting a thrust load disposed between a clutch hub of the clutch and the transmission case. An automatic transmission is proposed.

  According to the invention described in claim 2, in addition to the structure of claim 1, the first member extending radially inward from the ring gear of the planetary gear mechanism C, the clutch hub of the clutch, and the first A second member that is fixed to the input shaft so as not to rotate relative to the input shaft and extends radially outward; and a bearing C that can support a thrust load disposed between the first member and the second member. An automatic transmission characterized by this is proposed.

  According to the third aspect of the present invention, in addition to the configuration of the second aspect, the sun gear of the planetary gear mechanism B transmits the thrust load to the bearing A from the load transmission surface. An automatic transmission characterized in that a distance from C to a load receiving surface that receives a thrust load is larger than a total value of axial lengths of a plurality of components arranged between the load transmitting surface and the load receiving surface. Is proposed.

  According to the invention described in claim 4, in addition to the structure of any one of claims 1 to 3, the planetary gear mechanism C is fitted to the outer periphery of the input shaft so as to be relatively rotatable. A connecting member that connects the carrier of the planetary gear mechanism A to the carrier of the planetary gear mechanism A, and an outer peripheral spline that is fitted to an inner peripheral spline of the carrier of the planetary gear mechanism A is formed at one axial end of the connecting member. An automatic transmission is proposed in which an inner diameter of the inner peripheral spline is larger than an outer diameter of a sun gear of the planetary gear mechanism A.

  The second planetary gear mechanism PGSb of the embodiment corresponds to the planetary gear mechanism B of the present invention, and the third planetary gear mechanism PGSc of the embodiment corresponds to the planetary gear mechanism A of the present invention. The fourth planetary gear mechanism PGSd corresponds to the planetary gear mechanism C of the present invention, the third clutch C3 of the embodiment corresponds to the clutch of the present invention, and the seventh thrust bearing T7 and the eighth thrust bearing T8 of the embodiment. Corresponds to the bearing A capable of supporting the thrust load of the present invention, and the twelfth thrust bearing T12 of the embodiment corresponds to the bearing C capable of supporting the thrust load of the present invention, and the thirteenth thrust bearing T13 of the embodiment. Corresponds to the bearing B capable of supporting the thrust load according to the present invention, and the clutch drum 39 according to the embodiment corresponds to the first member according to the present invention. Strike plate 49 corresponds to the second member of the present invention.

  According to the configuration of the first aspect, the automatic transmission is disposed adjacent to the planetary gear mechanism A having the sun gear that cannot rotate relative to the input shaft, and the planetary gear mechanism A, and each element rotates relative to the planetary gear mechanism A. The planetary gear mechanism B that is possible, the planetary gear mechanism C that is disposed on the opposite side of the planetary gear mechanism B across the planetary gear mechanism A, and the ring gear of the planetary gear mechanism C are coupled to the input shaft. With possible clutches. Since the bearing A capable of supporting the thrust load disposed between the sun gear of the planetary gear mechanism A and the sun gear of the planetary gear mechanism B is provided, the thrust load acting on the sun gear of the planetary gear mechanism B is transferred from the bearing A to the planetary gear mechanism A. By transmitting to the input shaft via the sun gear, it is avoided as much as possible that the thrust load is transmitted from the planetary gear mechanism A to the input shaft via the planetary gear mechanism C, and the planetary gear mechanism A and the planetary gear mechanism C are avoided. It is possible to reduce the friction loss of a plurality of bearings interposed between these elements. Moreover, since the bearing B arranged between the clutch hub of the clutch and the transmission case capable of supporting the thrust load is provided, the thrust load transmitted to the input shaft can be supported by the transmission case via the bearing B. .

  According to the second aspect of the present invention, the first member extending radially inward from the ring gear of the planetary gear mechanism C and the clutch hub of the clutch and the first member are fixed to the input shaft so as not to rotate relative to each other. Since the second member extending outward and the bearing C capable of supporting the thrust load disposed between the first member and the second member are provided, the thrust load acting on the ring gear of the planetary gear mechanism C is directly applied to the input shaft. Can be transmitted and supported.

  According to the third aspect of the present invention, the distance from the load transmission surface where the sun gear of the planetary gear mechanism B transmits the thrust load to the bearing A to the load receiving surface where the second member receives the thrust load from the bearing C is the load transmission. A situation in which the thrust load of the sun gear of the planetary gear mechanism B is transmitted to the input shaft through the plurality of components because it is greater than the total axial length of the plurality of components disposed between the surface and the load receiving surface. Thus, friction can be reliably reduced.

  According to the fourth aspect of the present invention, the coupling member is fitted to the outer periphery of the input shaft so as to be relatively rotatable, and couples the carrier of the planetary gear mechanism C and the carrier of the planetary gear mechanism A, the axial direction of the coupling member One end is formed with an outer peripheral spline that fits into the inner peripheral spline of the carrier of the planetary gear mechanism A. Since the inner diameter of the inner peripheral spline is larger than the outer diameter of the sun gear of the planetary gear mechanism A, a connecting member is assembled in advance. When the planetary gear mechanism A is fitted to the outer periphery of the input shaft and the inner peripheral spline of the carrier of the planetary gear mechanism A is coupled to the outer peripheral spline of the connecting member, the inner peripheral spline of the carrier interferes with the sun gear of the planetary gear mechanism A. As a result, it is possible to avoid the situation where the assembly becomes impossible.

Skeleton diagram of automatic transmission. FIG. 2 is a detailed view of part 2 of FIG. 1. FIG. 3 is a detailed view of part 3 of FIG. The alignment chart of a 1st-4th planetary gear mechanism. Clutch and brake engagement table. Action | operation explanatory drawing at the time of the assembly of an automatic transmission. The table | surface which shows the direction of the torque of the sun gear of the 1st-4th planetary gear mechanism in each gear stage. The table | surface which shows the magnitude | size of the differential rotation of the 1st-13th thrust bearing in each gear stage. The table | surface which shows the magnitude | size of the thrust load of the 1st-13th thrust bearing in each gear stage. The table | surface which shows ratio of the thrust load of the 1st-4th planetary gear mechanism in each gear stage. The figure which shows the transmission path | route of thrust load F2, F3. The figure which shows the transmission path | route of thrust load F1, F2, F3. The figure which shows the transmission path | route of the thrust load F1, F2, F3, F4 in 6th speed, 7th speed, 9th speed, and 10th speed. The figure which shows the comparative example corresponding to FIG. The figure which shows the transmission path | route of the thrust load F1, F2, F3, F4 in the 3rd speed and 4th speed gear stage.

  Hereinafter, embodiments of the present invention will be described with reference to FIGS.

  First, the structure of the automatic transmission T having 10 forward speeds and 1 reverse speed will be described with reference to the skeleton diagram of FIG. 1 and the detailed views of FIG. 2 and FIG.

  The crankshaft 11 of the engine E is connected to the input shaft 12 of the automatic transmission T via a torque converter TC. On the outer periphery of the input shaft 12, the first planetary gear mechanism PGSa, the second planetary gear mechanism PGSb, the third planetary gear mechanism PGSc are arranged from the side closer to the engine E (right side in the figure) to the side farther (left side in the figure). The fourth planetary gear mechanism PGSd is arranged in order. In addition, in order to establish the respective gear positions by switching the coupling relationship of the elements of the first to fourth planetary gear mechanisms PGSa, PGSb, PGSc, and PGSd, the first clutch C1, the second clutch C2, the third clutch C3, 1 brake B1, 2nd brake B2, 3rd brake B3, and 4th brake B4 are provided. The second brake B2 is a two-way clutch, and the direction of relative rotation with which the second brake B2 is engaged can be arbitrarily switched. Other clutches and brakes are of a wet multi-plate type.

  The single-pinion type first planetary gear mechanism PGSa includes a first sun gear Sa, a first carrier Ca, a first ring gear Ra, and a plurality of first pinions Pa, and is rotatably supported by the first carrier Ca. The first pinions Pa ... mesh with the first sun gear Sa and the first ring gear Ra at the same time.

  The first sun gear Sa can be coupled to the transmission case 13 via the first brake B1. That is, the first brake B1 includes a brake hub 14 that is integral with the first sun gear Sa, a plurality of friction engagement elements 15 that are disposed between the brake hub 14 and the transmission case 13, and an axial slide on the transmission case 13. When the brake piston 16 is driven by hydraulic pressure supplied to the oil chamber 17 and the friction engagement elements 15 are engaged with each other, the first sun gear Sa is provided. Coupled to the transmission case 13.

  The transmission case 13 is actually composed of a plurality of members, but is illustrated as one member for convenience in the drawings. In the present specification, the stator shaft 19 fixed to the torque converter case 18 is also a part of the transmission case 13.

  The connecting member 20 extending radially outward from the first carrier Ca can be coupled to the transmission case 13 via the second brake B2. The first carrier Ca can be coupled to the input shaft 12 via the first clutch C1. That is, the first clutch C1 includes a clutch hub 22 connected to the radially inner end of the first carrier Ca through the connecting member 21, a clutch drum 23 fixed to the input shaft 12, and the clutch hub 22 and the clutch drum. Are provided between a plurality of friction engagement elements 24... And a clutch piston 25 slidably disposed inside the clutch drum 23. The clutch piston 25 is hydraulically supplied to the oil chamber 26. When the frictional engagement elements 24 are engaged with each other, the first carrier Ca is coupled to the input shaft 12.

  The first ring gear Ra is connected to a third carrier Cc of a third planetary gear mechanism PGSc, which will be described later, via a connecting member 27.

  The single pinion type second planetary gear mechanism PGSb includes a second sun gear Sb, a second carrier Cb, a second ring gear Rb, and a plurality of second pinions Pb... And is rotatably supported by the second carrier Cb. The second pinions Pb... Mesh with the second sun gear Sb and the second ring gear Rb at the same time.

  The second sun gear Sb is connected to a third ring gear Rc of a third planetary gear mechanism PGSc, which will be described later, via a connecting member 28. The second carrier Cb is integrally connected to the first carrier Ca of the first planetary gear mechanism PGSa via the connecting member 20, and can be coupled to the transmission case 13 via the second brake B2. An output gear 29 is formed integrally with the second ring gear Rb, and the output gear 29 is rotatably supported by the transmission case 13 via a pair of ball bearings 30 and 30.

  The single pinion type third planetary gear mechanism PGSc includes a third sun gear Sc, a third carrier Cc, a third ring gear Rc, and a plurality of third pinions Pc, and is rotatably supported by the third carrier Cc. The third pinions Pc... Mesh with the third sun gear Sc and the third ring gear Rc simultaneously.

  Third sun gear Sc is integrally coupled to input shaft 12. The third carrier Cc is connected to the first ring gear Ra of the first planetary gear mechanism PGSa via the connecting member 27. The third ring gear Rc can be coupled to a clutch drum 31 connected to a fourth sun gear Sd of a fourth planetary gear mechanism PGSd, which will be described later, via a second clutch C2. That is, the second clutch C2 includes a plurality of friction engagement elements 32 arranged between the third ring gear Rc and the clutch drum 31, and a clutch piston 33 slidably arranged inside the clutch drum 31. When the clutch piston 33 is driven by the hydraulic pressure supplied to the oil chamber 34 and the friction engagement elements 32 are engaged with each other, the third ring gear Rc of the third planetary gear mechanism PGSc becomes the fourth planetary gear mechanism. It is coupled to the fourth sun gear Sd of PGSd.

  The single pinion type fourth planetary gear mechanism PGSd includes a fourth sun gear Sd, a fourth carrier Cd, a fourth ring gear Rd, and a plurality of fourth pinions Pd, and is rotatably supported by the fourth carrier Cd. The fourth pinions Pd... Mesh with the fourth sun gear Sd and the fourth ring gear Rd simultaneously.

  The clutch drum 31 connected to the fourth sun gear Sd can be coupled to the transmission case 13 via the third brake B3. That is, the third brake B3 includes a plurality of friction engagement elements 35 arranged between the clutch drum 31 and the transmission case 13, and a brake piston 36 slidably arranged inside the transmission case 13. The fourth sun gear Sd is coupled to the transmission case 13 when the brake piston 36 is driven by the hydraulic pressure supplied to the oil chamber 37 to engage the friction engagement elements 35.

  The fourth carrier Cd is connected to the third carrier Cc of the third planetary gear mechanism PGSc via the connecting member 38. The fourth ring gear Rd can be coupled to the input shaft 12 via the third clutch C3. That is, the third clutch C3 includes a clutch drum 39 integrally connected to the fourth ring gear Rd, a clutch hub 40 integrally connected to the input shaft 12, and a plurality of clutches disposed between the clutch drum 39 and the clutch hub 40. Friction engagement elements 41... And a clutch piston 42 slidably disposed inside the clutch drum 39. The clutch piston 42 is driven by the hydraulic pressure supplied to the oil chamber 43 for friction engagement. When the elements 41 are engaged with each other, the fourth ring gear Rd is coupled to the input shaft 12.

  Further, the fourth ring gear Rd can be coupled to the transmission case 13 via the fourth brake B4. That is, the fourth brake B4 includes a plurality of friction engagement elements 44 disposed between the clutch drum 39 and the transmission case 13, and a brake piston 45 slidably disposed inside the transmission case 13. When the brake piston 45 is driven by the hydraulic pressure supplied to the oil chamber 46 and the friction engagement elements 44 are engaged with each other, the fourth ring gear Rd is coupled to the transmission case 13.

The shaft end of the input shaft 12 on the engine E side is directly supported by a stator shaft 19 constituting a part of the transmission case 13 via a ball bearing 47, and the shaft end of the input shaft 12 opposite to the engine E is located there. The clutch hub 40 fixed to is supported by the transmission case 13 via a ball bearing 48.

  Each gear of the first planetary gear mechanism PGSa, the second planetary gear mechanism PGSb, the third planetary gear mechanism PGSc, and the fourth planetary gear mechanism PGSd is composed of a helical gear to reduce the variation in transmission torque, Since the axial thrust force acts by the force, 13 thrust bearings are arranged between the relatively rotating members.

  The first thrust bearing T1 is disposed between the stator shaft 19 constituting a part of the transmission case 13 and the clutch drum 23 of the first clutch C1. The second thrust bearing T2 is disposed between the clutch drum 23 and the clutch hub 22 of the first clutch C1. The third thrust bearing T3 is disposed between the clutch hub 22 and the first sun gear Sa of the first planetary gear mechanism PGSa. The fourth thrust bearing T4 is disposed between the first sun gear Sa and the first carrier Ca of the first planetary gear mechanism PGSa. The fifth thrust bearing T5 is disposed between the coupling member 27 connected to the first ring gear Ra of the first planetary gear mechanism PGSa and the second sun gear Sb of the second planetary gear mechanism PGSb. The sixth thrust bearing T6 is disposed between the connecting member 27 and the first carrier Ca of the first planetary gear mechanism PGSa.

  The seventh thrust bearing T7 is disposed between the second sun gear Sb of the second planetary gear mechanism PGSb and the third carrier Cc of the third planetary gear mechanism PGSc. The eighth thrust bearing T8 is disposed between the third carrier Cc and the third sun gear Sc of the third planetary gear mechanism PGSc. The ninth thrust bearing T9 is disposed between the third carrier Cc and the clutch drum 31 of the second clutch C2. The tenth thrust bearing T10 is disposed between the fourth sun gear Sd and the fourth carrier Cd of the fourth planetary gear mechanism PGSd. The eleventh thrust bearing T11 is disposed between the fourth carrier Cd and the clutch drum 39 of the third clutch C3. The twelfth thrust bearing T12 is disposed between the clutch drum 39 and a thrust plate 49 fixed to the input shaft 12. The thirteenth thrust bearing T13 is disposed between the clutch hub 40 of the third clutch C3 and the transmission case 13.

  A shim 50 (see FIG. 3) is disposed between the fourth sun gear Sd of the fourth planetary gear mechanism PGSd and the clutch drum 31 of the second clutch C2. The first carrier Ca of the first planetary gear mechanism PGSa and the connecting member 21 extending to the left in the drawing from the clutch hub 22 of the first clutch C1 are abutted by the abutting portion 51 (see FIG. 2). Between the first thrust bearing T1 and the clutch drum 23 of the first clutch C1, a shim 52 (see FIG. 2) for adjusting the dimensional error of each component stacked on the input shaft 12 is disposed. .

The distance from the load transmitting surface P1 where the second sun gear Sb of the second planetary gear mechanism PGSb transmits the thrust load to the seventh thrust bearing T7 and the load receiving surface P2 where the thrust plate 49 receives the thrust load from the twelfth thrust bearing T12. L1 is a plurality of components arranged between the load transmission surface P1 and the load receiving surface P2, that is, the seventh thrust bearing T7, the third carrier Cc of the third planetary gear mechanism PGSc, the ninth thrust bearing T9, and the second clutch C2. Clutch drum 31, shim 50, fourth sun gear Sd of fourth planetary gear mechanism PGSd, tenth thrust bearing T10, fourth carrier Cd of fourth planetary gear mechanism PGSd, eleventh thrust bearing T11, clutch of third clutch C3 Axial direction of drum 39 and 12th thrust bearing T12 It is set larger than the sum of the length (see FIG. 3).

  An outer peripheral spline SPo formed on the outer peripheral surface of the right end in the drawing of the connecting member 38 connected to the fourth carrier Cd of the fourth planetary gear mechanism PGSd, and an inner periphery of the third carrier Cc of the third planetary gear mechanism PGSc in the left end of the drawing. The inner peripheral spline SPi formed on the surface meshes (see FIG. 6). The inner diameter r1 of the inner peripheral spline SPi of the third carrier Cc is set larger than the outer diameter r2 of the third sun gear Sc of the third planetary gear mechanism PGSc.

  The first carrier Ca of the first planetary gear mechanism PGSa receives a load from the load transmission surface P3 that transmits the thrust load to the fourth thrust bearing T4, and the clutch hub 22 of the first clutch C1 receives the thrust load from the third thrust bearing T3. The distance L2 to the surface P4 is set larger than the total value of the axial lengths of the fourth thrust bearing T4, the first sun gear Sa of the first planetary gear mechanism PGSa, and the third thrust bearing T3 (see FIG. 2). .

  FIG. 4 is a collinear diagram of the first to fourth planetary gear mechanisms PGSa, PGSb, PGSc, and PGSd. This corresponds to PGSa and the second planetary gear mechanism PGSb.

  For example, the gear ratio of the first planetary gear mechanism PGSa shown in the third stage from the top is h, the distance between the first sun gear Sa and the first carrier Ca, and the distance between the first carrier Ca and the first ring gear Ra. Is set to h: 1. Similarly, the gear ratio of the second planetary gear mechanism PGSb is i, the gear ratio of the third planetary gear mechanism PGSc is j, and the gear ratio of the fourth planetary gear mechanism PGSd is k.

  Of the two parallel horizontal lines in each collinear diagram, the upper horizontal line means that the rotational speed is “1” (the same rotational speed as the input shaft 12), and the lower horizontal line indicates that the rotational speed is “0”. ”(Stop). A speed line indicated by a broken line indicates that, among the first to fourth planetary gear mechanisms PGSa, PGSb, PGSc, and PGSd, other planetary gear mechanisms are idled following the planetary gear mechanism that transmits power.

  FIG. 7 shows the directions of torque acting on the first to fourth sun gears Sa, Sb, Sc, Sd of the first planetary gear mechanism PGSa, the second planetary gear mechanism PGSb, the third planetary gear mechanism PGSc, and the fourth planetary gear mechanism PGSd. “+” Indicates that the torque direction of the engine E is the same, and “−” indicates that the torque direction of the engine E is opposite. The direction of the thrust load acting on the first to fourth sun gears Sa, Sb, Sc, Sd made of helical gears is determined according to the direction of torque.

As shown in FIGS. 11 to 15, the first sun gear Sa of the first planetary gear mechanism PGSa is subjected to a leftward thrust load in the drawings at the first to fifth gears, and at the other gears, the thrust load is applied. The load does not work. On the second sun gear Sb of the second planetary gear mechanism PGSb, a leftward thrust load in the drawing acts on all the forward gears, and a rightward thrust load on the reverse gears acts. On the third sun gear Sc of the third planetary gear mechanism PGSc, a leftward thrust load acts in the forward shift speed except the 8-speed shift speed, and a rightward thrust load acts in the reverse speed speed. The fourth sun gear Sd of the fourth planetary gear mechanism PGSd is subjected to a leftward thrust load in the figure at the sixth speed, the seventh speed, the ninth speed and the tenth speed, and the third speed, A rightward thrust load acts in the speed shift stage and the reverse shift stage.

  FIG. 8 shows an example of differential rotation of the first to thirteenth thrust bearings T <b> 1 to T <b> 13 at each gear stage, and 1,000 rpm corresponds to differential rotation between the input shaft 12 and the transmission case 13. For example, the differential rotation between the first thrust bearing T1 and the thirteenth thrust bearing T13 disposed between the input shaft 12 and the transmission case 13 is 1,000 rpm.

  FIG. 9 shows an example of the thrust load acting on the first to thirteenth thrust bearings T1 to T13 at each shift stage. Since the thrust load acting on the first to thirteenth thrust bearings T1 to T13 varies depending on the torsion angle of the gear, an example thereof is shown in FIG.

  Since the sun gear and the ring gear of the single pinion type planetary gear mechanism mesh with the common pinion, the thrust loads received by the sun gear and the ring gear are the same and opposite to each other. FIG. 10 shows the thrust load F1 of the first sun gear Sa and the first ring gear Ra of the first planetary gear mechanism PGSa, the thrust load F2 of the second sun gear Sb and the second ring gear Rb of the second planetary gear mechanism PGSb, and the third planetary gear. The thrust load F3 of the third sun gear Sc and the third ring gear Rc of the gear mechanism PGSc, and the thrust load F4 of the fourth sun gear Sd and the fourth ring gear Rd of the fourth planetary gear mechanism PGSd are shown for each shift stage. . In addition, the numerical value of F1-F4 in FIG. 10 has shown the relative magnitude | size between them.

  Next, an embodiment of the present invention having the above configuration will be described.

  First, the torque flow of each gear stage will be described based on the skeleton diagram of FIG. 1, the alignment chart of FIG. 4, and the engagement table of FIG. In the engagement table of FIG. 5, the circles indicate that the clutch or brake is in an engaged state. Further, since the second brake B2 is constituted by a two-way clutch, the second brake B2 is switched between the forward rotation prevention state F and the reverse rotation prevention state R. Underlined F and R indicate that the rotational speeds of the first carrier Ca and the second carrier Cb become “0” by the action of the second brake B2.

  When the first gear is established, the first brake B1 is engaged, the second brake B2 is set in the reverse rotation prevention state, and the third brake B3 is engaged. By setting the second brake B2 to the reverse rotation prevention state, the reverse rotation of the first carrier Ca and the second carrier Cb is prevented. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. As a result, the three elements of the first planetary gear mechanism PGSa enter a locked state in which relative rotation is impossible, and the rotational speeds of the first ring gear Ra, the third carrier Cc, and the fourth carrier Cd also become “0”. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “1st” shown in FIG. 4, and the first gear is established.

  It is not necessary to engage the third brake B3 in order to establish the first gear, but if the third brake B3 is engaged in advance while the first gear is established, the first gear is changed. The transition to the second gear can be smoothly performed. If it is necessary to operate the engine brake at the first gear, the second brake B2 may be switched from the reverse rotation prevention state to the forward rotation prevention state.

  When the second gear is established, the first brake B1 is engaged, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, and the second clutch C2 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the second clutch C2, the rotation speeds of the third ring gear Rc and the second sun gear Sb become “0”, which is the same as the rotation speed of the fourth sun gear Sd. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “2nd” shown in FIG. 4, and the second gear is established.

  When the third speed is established, the first brake B1 is engaged, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, and the third clutch C3 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the third clutch C <b> 3, the rotational speed of the fourth ring gear Rd becomes “1”, which is the same speed as the rotational speed of the third sun gear Sc connected to the input shaft 12. In this way, the rotational speed of the fourth sun gear Sd becomes “0” and the rotational speed of the fourth ring gear Rd becomes “1”, so the rotational speed of the fourth carrier Cd, that is, the fourth carrier Cd, the third The rotational speeds of the carrier Cc and the first ring gear Ra are k / (k + 1). As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “3rd” shown in FIG. 4, and the third gear is established.

  When the 4th speed is established, the first brake B1 is engaged, the second brake B2 is set in the reverse rotation prevention state, the second clutch C2 is engaged, and the third clutch C3 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. By engaging the second clutch C2, the fourth sun gear Sd, the third ring gear Rc, and the second sun gear Sb rotate at the same speed. As a result, the third carrier Cc and the fourth carrier Cd are connected between the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd, and the third ring gear Rc and the fourth sun gear Sd are connected. In the fourth speed gear stage that engages the second clutch C2, the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd can draw one collinear diagram composed of four elements.

By engaging the third clutch C3, the rotation speed of the fourth ring gear Rd becomes “1”, which is the same as the rotation speed of the third sun gear Sc, and the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd. rotational speed of the two elements of the four element becomes "1" in the same speed that make up and. As a result, the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd are in a locked state where the respective elements cannot rotate relative to each other, and the rotational speeds of all the elements of the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd are It becomes “1”. Then, the rotational speeds of the first carrier Ca and the second carrier Cb become h / (h + 1), and the rotational speed of the second ring gear Rb to which the output gear 29 is connected becomes “4th” shown in FIG. A stage is established.

  When the fifth gear is established, the first brake B1 is engaged, the second brake B2 is in the reverse rotation prevention state, the first clutch C1 is engaged, and the third clutch C3 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the first brake B1, the rotation speed of the first sun gear Sa becomes “0”. By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “5th” shown in FIG. 4, and the fifth gear is established.

  It is not necessary to engage the third clutch C3 when establishing the fifth gear, but the third clutch C3 needs to be engaged at the fourth and sixth gears adjacent to the fifth gear. Therefore, by engaging the third clutch C3 while establishing the fifth gear, the shift change to the fourth gear or the sixth gear can be performed smoothly.

When the sixth speed is established, the second brake B2 is set in the reverse rotation prevention state, the first clutch C1, the second clutch C2, and the third clutch C3 are engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the second clutch C2 and the third clutch C3, the third planetary gear mechanism PGSc and the fourth planetary gear mechanism PGSd are brought into the locked state as described in the fourth speed, and the third ring gear Rc. The rotational speed of the second sun gear Sb is “1”. By engaging the first clutch C1 , the rotation speeds of the first carrier Ca and the second carrier Cb become “1”.

  Thereby, in the second planetary gear mechanism PGSb, the rotation speeds of the second carrier Cb and the second sun gear Sb become “1” of the same speed, and each element is in a locked state in which the relative rotation is impossible. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “1” of “6th” shown in FIG. 4, and the sixth speed gear stage is established.

  When the seventh speed is established, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, the first clutch C1 is engaged, and the third clutch C3 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the third clutch C3, the rotational speed of the fourth ring gear Rd becomes “1”, which is the same as the rotational speed of the third sun gear Sc, and the third carrier Cc, the first ring gear Ra, and the fourth carrier Cd. Becomes the rotation speed k / (k + 1). By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc connected to the input shaft 12. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “7th” shown in FIG. 4, and the seventh speed gear stage is established.

  At the time of establishment of the eighth gear, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, the first clutch C1 is engaged, and the second clutch C2 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the second clutch C2, the rotational speeds of the third ring gear Rc and the second sun gear Sb become “0”, which is the same speed as the rotational speed of the fourth sun gear Sd. By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “8th” shown in FIG. 4, and the 8-speed gear stage is established.

  When the ninth gear is established, the second brake B2 is set in the reverse rotation prevention state, the third brake B3 is engaged, the fourth brake B4 is engaged, and the first clutch C1 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the third brake B3, the rotational speed of the fourth sun gear Sd becomes “0”. By engaging the fourth brake B4, the rotational speed of the fourth ring gear Rd is also "0". As a result, the fourth planetary gear mechanism PGSd is locked, and the rotational speeds of the fourth carrier Cd, the third carrier Cc, and the first ring gear Ra are also “0”.

  By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “9th” shown in FIG. 4, and the ninth gear is established.

  When the tenth speed is established, the second brake B2 is set in the reverse rotation prevention state, the fourth brake B4 is engaged, the first clutch C1 is engaged, and the second clutch C2 is engaged. The first carrier Ca and the second carrier Cb are allowed to rotate forward by setting the second brake B2 to the reverse rotation preventing state. By engaging the second clutch C2, the third ring gear Rc and the second sun gear Sb rotate at the same speed as the fourth sun gear Sd. By engaging the fourth brake B4, the rotational speed of the fourth ring gear Rd becomes “0”. By engaging the first clutch C1, the rotational speeds of the first carrier Ca and the second carrier Cb become “1”, which is the same speed as the rotational speed of the third sun gear Sc. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “10th” shown in FIG.

  At the time of establishment of the reverse gear, the second brake B2 is set in the forward rotation blocking state, the third brake B3 is engaged, and the third clutch C3 is engaged. By engaging the third brake B3 and the third clutch C3, the rotational speeds of the third carrier Cc, the first ring gear Ra, and the fourth carrier Cd become k / (k + 1). By setting the second brake B2 in the forward rotation blocking state, the forward rotation of the first carrier Ca and the second carrier Cb is blocked, and the rotation speeds of the first carrier Ca and the second carrier Cb become “0”. As a result, the rotation speed of the second ring gear Rb to which the output gear 29 is connected becomes “Rvs” shown in FIG. 4, and the reverse gear position is established.

  Next, reduction of thrust bearing friction will be described.

The second sun gear Sb of the second planetary gear mechanism PGSb is subjected to a leftward thrust load F2 in the drawing at all forward speeds (see FIGS. 7 and 11). The thrust load F2 seventh thrust bearing T7, the third carrier Cc of the third planetary gear mechanism PGSc, is transmitted to the third input shaft 12 via the sun gear Sc of the eighth thrust bearing T8 and third planetary gear mechanism PGSc, input The transmission is transmitted from the shaft 12 to the transmission case 13 via the clutch hub 40 of the third clutch C3 and the thirteenth thrust bearing T13.

At that time, as shown in FIG. 3, the second sun gear Sb of the second planetary gear mechanism PGSb transmits the thrust load to the seventh thrust bearing T7, and the thrust plate 49 extends from the twelfth thrust bearing T12. The distance L1 to the load receiving surface P2 that receives the thrust load is the seventh thrust bearing T7 disposed between the load transmitting surface P1 and the load receiving surface P2, the third carrier Cc of the third planetary gear mechanism PGSc, and the ninth thrust bearing. T9, clutch drum 31 of the second clutch C2, shim 50, fourth sun gear Sd of the fourth planetary gear mechanism PGSd, tenth thrust bearing T10, fourth carrier Cd of the fourth planetary gear mechanism PGSd, eleventh thrust bearing T11, The clutch drum 39 and the twelfth thrust bearing T of the third clutch C3 2 is set to be larger than the total value of the axial lengths of 2, the thrust load F2 of the second sun gear Sb of the second planetary gear mechanism PGSb is transmitted to the transmission case 13 via the input shaft 12 through the shortest path. Will be. As a result, the thrust load F2 is prevented from passing through many parts such as the fourth planetary gear mechanism PGSd and the thrust bearings T9 to T12, and the friction can be reduced and the parts can be made thinner.

  Further, at the third speed gear stage and the fourth speed gear stage, a leftward thrust load F4 (see FIGS. 7 and 15) acts on the fourth ring gear Rd of the fourth planetary gear mechanism PGSd. The transmission can be transmitted from the 4-ring gear Rd to the input shaft 12 via the clutch drum 39, the twelfth thrust bearing T12 and the thrust plate 49, and supported by the transmission case 13 via the thirteenth thrust bearing T13.

  FIG. 6 shows a procedure for determining the thickness of the shim 50 (see FIG. 3). First, the third clutch C3, the fourth brake B4, the third brake B3, the fourth planetary gear mechanism PGSd, and the like are assembled on the outer periphery of the input shaft 12. In this state, from the surface of the fourth planetary gear mechanism PGSd that contacts the shim 50 of the fourth sun gear Sd to the right end of the third sun gear Sc of the third planetary gear mechanism PGSc that is pre-splined to the input shaft 12 in the drawing. Measure the distance X.

  Next, the distance Y from the left end (surface contacting the shim 50) of the clutch drum 31 of the second clutch C2 assembled in advance as a subassembly to the right end (surface contacting the ninth thrust bearing T9) in the drawing is measured. To do. The third carrier Cc, the third pinion Pc, the eighth thrust bearing T8, and the ninth thrust bearing T9 are preassembled as subassemblies, and the left end of the ninth thrust bearing T9 in the figure and the left end of the eighth thrust bearing T8 in the figure. The distance Z to is measured. Then, a value slightly smaller than X− (Y + Z) is set as the thickness of the shim 50, and the shim 50 corresponding to the above-described thickness may be selected from a plurality of types of shim 50 prepared in advance. .

  As described above, when the pre-assembled sub-assembly of the third carrier Cc is assembled so as to be fitted to the outer periphery of the input shaft 12, the inner peripheral spline SPi of the third carrier Cc extends the outer periphery of the third sun gear Sc. Although passing from the right side to the left side in the drawing, the radius r1 of the inner peripheral spline SPi of the third carrier Cc is set to be larger than the radius r2 of the third sun gear Sc, so that the third carrier Cc does not interfere with the third sun gear Sc. A subassembly of the planetary gear mechanism PGSc can be assembled.

As described above, the clutch drum 31 of the second clutch C 2 capable of coupling the third ring gear Rc of the third planetary gear mechanism PGSc and the fourth sun gear Sd of the fourth planetary gear mechanism PGSd and the fourth planetary gear mechanism PGSd Since the shim 50 is disposed between the fourth sun gear Sd and the shim 50, the relative distance between the elements of the fourth planetary gear mechanism PGSd or the distance between the elements of the second clutch C2 may be affected. In addition, the shim 50 can be arranged by utilizing the dead space between the fourth planetary gear mechanism PGSd and the second clutch C2, so that an increase in the axial dimension of the automatic transmission T is minimized. Can do.

As shown in FIG. 2, the end portion in the drawing of the left side of the first carrier Ca of the first planetary gear mechanism PGSa, and the end portion in the drawing left of the connecting member 21 extending from the clutch hub 22 of the first clutch C1 Is abutted at the butting portion 51, so that the rightward thrust load acting on the second sun gear Sb of the second planetary gear mechanism PGSb when the vehicle is traveling at a reduced speed (when the engine is braked) or when the vehicle is traveling backward is: Fifth thrust bearing T5 → connecting member 27 → sixth thrust bearing T6 → first carrier Ca of first planetary gear mechanism PGSa → abutting portion 51 → connecting member 21 → clutch hub 22 of the first clutch C1 → second thrust bearing Stator shaft 19 (transmission case 1) in the path of T2 → clutch drum 23 of first clutch C1, shim 52 → first thrust bearing T1. ) Is transmitted to the.

At this time, assuming that the abutment portion 51 does not exist, the load input from the sixth thrust bearing T6 to the first carrier Ca of the first planetary gear mechanism PGSa does not pass through the connecting member 21, but the fourth thrust bearing. The first planetary gear mechanism that rotates at a high speed is transmitted from the clutch hub 22 of the first clutch C1 to the second thrust bearing T2 through the path of T4 → first sun gear Sa of the first planetary gear mechanism PGSa → third thrust bearing T3. Large friction is generated in the fourth thrust bearing T4 and the third thrust bearing T3 on both sides of the first sun gear Sa of the PGSa. Moreover, when the thrust load passes through the fourth thrust bearing T4 and the third thrust bearing T3, the thickness type of the shim 52 increases in order to compensate for the dimensional tolerance of the fourth thrust bearing T4 and the third thrust bearing T3. There is a problem.

On the other hand, according to the present embodiment, by providing the abutment portion 51, the thrust load bypasses the fourth thrust bearing T4 and the third thrust bearing T3 that sandwich the first sun gear Sa of the first planetary gear mechanism PGSa. Therefore, not only the friction of the fourth thrust bearing T4 and the third thrust bearing T3 is reduced, but it is not necessary to consider these dimensional tolerances, and the thickness type of the shim 52 can be reduced.

  At this time, the first carrier Ca of the first planetary gear mechanism PGSa transmits a thrust load from the third thrust bearing T3 to the clutch hub 22 of the first clutch C1 from the load transmission surface P3 that transmits the thrust load to the fourth thrust bearing T4. The distance L2 to the load receiving surface P4 to be received is larger than the total value of the axial lengths of the fourth thrust bearing T4, the first sun gear Sa of the first planetary gear mechanism PGSa, and the third thrust bearing T3. To avoid the situation where the rightward thrust load input from the gear mechanism PGSb is transmitted to the transmission case 13 via the fourth thrust bearing T4 and the third thrust bearing T3, the friction can be reduced reliably. it can.

  The end portion of the first carrier Ca of the first planetary gear mechanism PGSa provided with the abutting portion 51 also serves as a holding portion for the sixth thrust bearing T6.

  Next, the setting of the direction of the helical angle of the helical gears of the first planetary gear mechanism PGSa, the second planetary gear mechanism PGSb, the third planetary gear mechanism PGSc, and the fourth planetary gear mechanism PGSd will be described.

  Since the second ring gear Rb of the second planetary gear mechanism PGSb is formed integrally with the output gear 29, the rightward thrust load F2 generated in the figure by the second ring gear Rb is transmitted via the output gear 29 and the ball bearings 30 and 30. And transmitted to the transmission case 13. Therefore, the leftward thrust load F2 generated by the second sun gear Sb of the second planetary gear mechanism PGSb directly cancels the rightward thrust load F2 generated by the second ring gear Rb of the second planetary gear mechanism PGSb. However, it is necessary to transmit and support the transmission case 13 by any route. At this time, if the thrust load F2 facing left in the drawing of the second sun gear Sb can be offset with the thrust load generated by another planetary gear mechanism, the thrust bearing existing in the path transmitted to the transmission case 13 will be described. Friction can be reduced.

  From this point of view, in the present embodiment, as shown in FIG. 11, the leftward thrust load F2 generated by the second sun gear Sb of the second planetary gear mechanism PGSb is applied to the third planetary gear mechanism PGSc adjacent thereto. The rightward thrust load F3 generated by the third ring gear Rc and the connecting member 28 immediately cancel each other, and only the thrust load F2-F3, which is the difference between the thrust load F3 and the third planetary gear mechanism PGSc, is offset. The transmission is transmitted to the transmission case 13 via the 3 sun gear Sc, the input shaft 12 and the thirteenth thrust bearing T13.

  As is apparent from FIG. 10, in this embodiment, the thrust load F2 facing left in the drawing of the second sun gear Sb is larger than the thrust load F3 facing right in the drawing of the third ring gear Rc in all forward speeds. The thrust load F2-F3 that is the difference is always leftward in the figure. However, the above-described relationship of F2> F3 does not always hold, and the gear twist angle may be set so that F2 <F3 at a low speed gear (for example, the first gear or the second gear). . This is because, at low speed gears used for starting and running at low speed, the loss of power transmission is originally large, so that the effect of reducing friction obtained is small even if the relationship of F2> F3 is established.

  A leftward thrust load F2 of the second sun gear Sb in the figure is input to the seventh thrust bearing T7 and the eighth thrust bearing T8, and a part of the thrust load F2 is a rightward thrust load of the third ring gear Rc in the figure. By canceling with F3, the load on the seventh thrust bearing T7 and the eighth thrust bearing T8 can be reduced.

  In FIG. 11, it is assumed that the first planetary gear mechanism PGSa and the fourth planetary gear mechanism PGSd are unloaded.

  As shown in FIGS. 7 and 12, the thrust load F1 acts on the first sun gear Sa and the first ring gear Ra of the first planetary gear mechanism PGSa only in the first gear to the fifth gear, and the first The leftward thrust load F1 of the sun gear Sa in the drawing is equal to the rightward thrust load F1 of the first ring gear Ra in the drawing. Then, the leftward thrust load F1 of the first sun gear Sa in the drawing is transmitted to the first ring gear Ra through the path of the fourth thrust bearing T4 → the first carrier Ca → the sixth thrust bearing T6 → the connecting member 27, and the first ring gear Ra. The thrust load F1 is completely offset with the rightward thrust load F1, so that the thrust load F1 does not affect other planetary gear mechanisms.

  The direction of the thrust load acting on the sun gear of the planetary gear mechanism is determined by the rotation direction (torque direction) of the sun gear and the direction of the torsion angle of the tooth surface, but the direction of the torsion angle does not change. The direction of thrust load changes according to the change of. In the single pinion type planetary gear mechanism, the thrust load acting on the sun gear and the thrust load acting on the ring gear are directed in opposite directions.

  In the present embodiment, since the rotation direction (torque direction) of the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is as shown in FIG. 7, the fourth sun gear Sd and the fourth ring gear of the fourth planetary gear mechanism PGSd. The thrust load F4 of Rd is rightward in the drawing and leftward in the drawing at the third gear, the fourth gear, and the reverse gear, respectively, but the sixth gear, the seventh gear, the ninth gear, and the tenth gear. At the speed shift stage, they change to the left in the figure and to the right in the figure, respectively. By setting the direction of the thrust load F4 of the fourth sun gear Sd and the fourth ring gear Rd as described above, the following operational effects can be obtained.

  As shown in FIG. 8, in the 9th speed stage and the 10th speed stage, which are the high speed speed stages, there is a problem that the differential rotation of the 12th thrust bearing T12 reaches 1000 rpm at the maximum and the friction increases. By setting the direction of the thrust load F4 of Sd and the fourth ring gear Rd as described above, the fourth sun gear Sd and the fourth ring gear Rd of the ninth speed gear and the tenth speed gear as shown in FIG. The thrust load F4 is canceled by the tenth thrust bearing T10 and the eleventh thrust bearing T11, and is not transmitted to the twelfth thrust bearing T12 having a large differential rotation, and the friction of the twelfth thrust bearing T12 can be effectively reduced.

  By the way, at the sixth speed, the rightward thrust load F4 acting on the fourth sun gear Sd of the fourth planetary gear mechanism PGSd increases (see FIG. 10). Accordingly, assuming that the direction of the thrust load F4 acting on the fourth ring gear Rd and the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is set opposite to the embodiment described in FIG. 13, as shown in FIG. Furthermore, the left thrust load F2-F3 in the drawing and the right thrust load F4 in the drawing transmitted from the fourth sun gear Sd through the ninth thrust bearing T9 are input to the seventh thrust bearing T7. At the sixth speed, the rightward thrust load F4 in the figure increases, so the resultant force F2-F3-F4 has a negative value, and the seventh thrust bearing T7 may be urged to the right in the figure. .

  In such a situation, the thrust load for urging the seventh thrust bearing T7 in the right direction in the figure is the seventh thrust bearing T7 → the fifth thrust bearing T5 → the sixth thrust bearing T6 → the second thrust bearing T2 → the first thrust. This is transmitted to the stator shaft 19 (transmission case 13) through the path of the bearing T1, and the friction of the five thrust bearings T7 to T5, T2, and T1 existing in the path increases.

  From the above, at the sixth speed, the seventh speed, the ninth speed and the tenth speed, the thrust load F4 of the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is directed leftward in the figure. It is necessary to set the helix angle of the helical gear.

  By the way, as described above, the thrust load F4 of the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is directed leftward in the drawing at the sixth speed, the seventh speed, the ninth speed, and the tenth speed. When the helical gear torsion angle is set, the direction of the thrust load F4 of the fourth sun gear Sd of the fourth planetary gear mechanism PGSd is reversed at the third speed gear stage and the fourth speed gear stage and becomes rightward in the figure (see FIG. 7). As described in FIG. 14, the thrust load F4 of the fourth sun gear Sd may be transmitted to the five thrust bearings T7 to T5, T2, and T1.

  However, the rightward thrust load F4 of the fourth sun gear Sd in the drawing at the third and fourth gears is smaller than the rightward thrust load F4 of the fourth sun gear Sd in the sixth gear (see the figure). 10), the resultant force F2-F3-F4 does not become negative, and the thrust load F4 is transmitted to the stator shaft 19 (transmission case 13) via five thrust bearings T7 to T5, T2, T1. Can be avoided (see FIG. 15).

  Further, in the third gear and the fourth gear, the leftward thrust load F4 of the fourth ring gear Rd of the fourth planetary gear mechanism PGSd is transmitted to the twelfth thrust bearing T12, but the third gear and the fourth gear. Since the differential rotation of the twelfth thrust bearing T12 is zero at the shift speed (see FIG. 8), there is no problem that the friction of the twelfth thrust bearing T12 increases even if the thrust load F4 is input.

  The embodiments of the present invention have been described above, but various design changes can be made without departing from the scope of the present invention.

  For example, although the third sun gear Sc of the third planetary gear mechanism PGSc is splined to the input shaft 12 in the embodiment, the third sun gear Sc may be formed integrally with the input shaft 12.

In the embodiment, the second brake B2 is a two-way clutch, but it can be a multi- plate brake.

  In the embodiment, the thirteenth thrust bearing T13 transmits the thrust load input from the input shaft 12 to the transmission case 13 through the outer ring of the ball bearing 48. However, the thirteenth thrust bearing T13 is abolished and the ball The thrust load input from the input shaft 12 may be transmitted to the transmission case 13 by changing the bearing 48 to a ball bearing (for example, an angular ball bearing) that can also support a thrust load.

P GS b Second planetary gear mechanism (planetary gear mechanism B)
Sb Sun gear P GS c of planetary gear mechanism B Third planetary gear mechanism (planetary gear mechanism A)
Sc Sun gear Cc of planetary gear mechanism A Carrier P GS d of planetary gear mechanism A Fourth planetary gear mechanism (planetary gear mechanism C)
Cd Carrier Rd of planetary gear mechanism C Ring gear C3 of planetary gear mechanism C Third clutch (clutch)
T7 7th thrust bearing (bearing A capable of supporting thrust load)
T8 8th thrust bearing (bearing A capable of supporting thrust load)
T12 12th thrust bearing (bearing C capable of supporting thrust load)
T13 13th thrust bearing (bearing B capable of supporting thrust load)
SPi Planetary gear mechanism A carrier inner peripheral spline SPo Connecting member outer peripheral spline r1 Inner peripheral spline inner diameter r2 Planetary gear mechanism A sun gear outer diameter P1 Load transmission surface P2 Load receiving surface L1 From load transmitting surface to load receiving surface Distance 12 Input shaft 13 Transmission case 38 Connecting member 39 Clutch drum (first member)
40 Clutch hub 49 Thrust plate (second member)

Claims (4)

A plurality of planetary gear mechanisms arranged on the axis of the input shaft (12) are housed in a transmission case (13), and each element of the plurality of planetary gear mechanisms, the input shaft (12), and the transmission case (13) In an automatic transmission that establishes a predetermined gear position by controlling a coupling relationship between a plurality of clutches and a plurality of brakes,
A planetary gear mechanism A (PSGc) having a sun gear (Sc) that cannot rotate relative to the input shaft (12);
A planetary gear mechanism B (PSGb) disposed adjacent to the planetary gear mechanism A (PSGc) and capable of rotating relative to each other;
A planetary gear mechanism C (PSGd) that is disposed on the opposite side of the planetary gear mechanism B (PSGb) with the planetary gear mechanism A (PSGc) interposed therebetween, and each element can relatively rotate;
A clutch (C3) capable of coupling the ring gear (Rd) of the planetary gear mechanism C (PSGd) to the input shaft (12);
A bearing A (T7, T8) capable of supporting a thrust load disposed between a sun gear (Sc) of the planetary gear mechanism A (PSGc) and a sun gear (Sb) of the planetary gear mechanism B (PSGb);
An automatic transmission comprising a bearing B (T13) capable of supporting a thrust load disposed between a clutch hub (40) of the clutch (C3) and the transmission case (13). A first member (39) extending radially inward from the ring gear (Rd) of the planetary gear mechanism C (PSGd);
A second member (49) extending radially outwardly fixed to the input shaft (12) so as not to rotate relative to the input shaft (12) between the clutch hub (40) of the clutch (C3) and the first member (39);
The automatic transmission according to claim 1, further comprising a bearing (T12) capable of supporting a thrust load disposed between the first member (39) and the second member (49).   The sun gear (Sb) of the planetary gear mechanism B (PSGb) transmits the thrust load to the bearings A (T7, T8) from the load transmission surface (P1), and the second member (49) extends from the bearing C (T12). The distance (L1) to the load receiving surface (P2) that receives the thrust load is the sum of the axial lengths of a plurality of components arranged between the load transmitting surface (P1) and the load receiving surface (P2). The automatic transmission according to claim 2, wherein A connecting member that fits on the outer periphery of the input shaft (12) so as to be relatively rotatable and connects the carrier (Cd) of the planetary gear mechanism C (PSGd) and the carrier (Cc) of the planetary gear mechanism A (PSGc). (38)
An outer peripheral spline (SPo) that fits with an inner peripheral spline (SPi) of the carrier (Cc) of the planetary gear mechanism A (PSGc) is formed at one end of the connecting member (38) in the axial direction. The inner diameter (r1) of (SPi) is larger than the outer diameter (r2) of the sun gear (Sc) of the planetary gear mechanism A (PSGc), according to any one of claims 1 to 3. Automatic transmission as described.

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