JP6139267B2 - Automatic transmission - Google Patents

Description

  The present invention includes a plurality of planetary gear mechanisms having elements of a sun gear, a carrier and a ring gear, and a plurality of engaging means for coupling the elements of the plurality of planetary gear mechanisms to each other or to other members. The present invention relates to an automatic transmission that establishes a plurality of shift stages in response to engagement and disengagement of the plurality of engagement means.

  In such an automatic transmission, by combining four planetary gear mechanisms, three clutches, and four brakes, it is possible to establish nine or more forward gears and one reverse gear. This is known from US Pat.

JP 2012-154425 A

  By the way, the conventional one is configured such that the ring gear Rb of the second planetary gear mechanism PGS2 can be coupled to the transmission case 1 by the fourth brake B4 and can be coupled to the input shaft 2 by the second clutch C2. . Since the fourth brake B4 is disposed so as to surround the radially outer side of the ring gear Rb, there has been a problem that the outer diameter of the fourth brake B4 increases and the radial dimension of the automatic transmission increases. The second clutch C2 includes a clutch drum extending radially outward from the input shaft 2 and a clutch hub extending radially inward along the side portion of the second planetary gear mechanism PGS2 from the ring gear Rb. Therefore, there is a problem that the torque transmission path from the ring gear Rb to the input shaft 2 is bent and complicated, and the axial dimension of the automatic transmission is increased.

  The present invention has been made in view of the above circumstances, and a brake that couples a ring gear of a planetary gear mechanism of an automatic transmission to a transmission case and a clutch that couples the ring gear to an input shaft are rationally arranged, The purpose is to reduce the size of the automatic transmission.

To achieve the above object, according to the first aspect of the present invention, a plurality of planetary gear mechanisms having elements of a sun gear, a carrier, and a ring gear are coupled to each other of the elements of the plurality of planetary gear mechanisms. Or a plurality of planetary gears in an automatic transmission that includes a plurality of engagement means coupled to other members, and that establishes a plurality of shift stages according to engagement and disengagement of the plurality of engagement means. A multi-plate clutch for coupling one ring gear of the mechanism to an input shaft; and a multi-plate brake for coupling the ring gear to a transmission case; and the clutch includes a clutch hub coupled to the input shaft; A clutch drum having a cylindrical portion coupled to the ring gear, and a plurality of first friction engagement elements disposed between an outer peripheral portion of the clutch hub and an inner peripheral portion of the cylindrical portion, The rake includes a plurality of second friction engagement elements disposed between an outer peripheral portion of the cylindrical portion and an inner peripheral portion of the transmission case, and the first friction engagement element and the second friction engagement element are At least a portion overlaps in the axial direction, the clutch hub is coupled to the input shaft within the axial width of the plurality of first friction engagement elements, and the ring gear and the cylindrical portion couple them There is proposed an automatic transmission characterized in that they extend in opposite directions along the axial direction from the coupling portion.

  According to the invention described in claim 2, in addition to the configuration of claim 1, the first to fourth planetary gear mechanisms, the first to third clutches, and the first to fourth brakes are provided. The three elements of the first planetary gear mechanism are respectively designated as the first element, the second element, and the third element in order from the one side in the order corresponding to the gear ratio in the schematic diagram. The four elements, the fifth element and the sixth element are arranged in order from one side at intervals corresponding to the gear ratio in the schematic diagram, and the three elements of the third planetary gear mechanism are defined in the schematic diagram. The seventh element, the eighth element, and the ninth element are arranged in order from one side at intervals corresponding to the tooth number ratio, and the three elements of the fourth planetary gear mechanism are arranged at intervals corresponding to the tooth number ratio in the schematic diagram. 10th element, 11th element and 12th element from one side in order The seventh element is connected to the input shaft, the fourth element is connected to the output member, the eighth element, the eleventh element, and the third element are connected to form a first connector, The ninth element and the sixth element are connected to form a second connecting body, the second element and the fifth element are connected to form a third connecting body, and the first clutch is The seventh element can be coupled to the third coupling body, the second clutch can be coupled to the twelfth element and the second coupling body, and the third clutch can be coupled to the seventh element. The tenth element can be coupled, the first brake can couple the first element to the transmission case, and the second brake can couple the third connector to the transmission case. And the third brake couples the twelfth element to the transmission case. An ability, the fourth brake can be coupled to the first 10 elements in the transmission case, the automatic transmission, wherein the ring gear is the tenth element is proposed.

  The fourth ring gear Rd of the embodiment corresponds to the ring gear of the invention of claim 1 and also corresponds to the tenth element of the invention of claim 2, and the third clutch C3 of the embodiment corresponds to the invention of claim 1. Corresponding to the clutch, the fourth brake B4 of the embodiment corresponds to the brake of the invention of claim 1, the output gear 29 of the embodiment corresponds to the output member of the invention of claim 2, and the friction of the embodiment. The engagement element 41 corresponds to the first friction engagement element of the invention of claim 1, and the friction engagement element 44 of the embodiment corresponds to the second friction engagement element of the invention of claim 1.

  According to the configuration of the first aspect, the clutch for coupling the ring gear of the planetary gear mechanism to the input shaft includes a clutch hub coupled to the input shaft, a clutch drum having a cylindrical portion coupled to the ring gear, and an outer periphery of the clutch hub. And a plurality of first frictional engagement elements disposed between the inner peripheral portion of the cylindrical portion and a brake for coupling the ring gear of the planetary gear mechanism to the transmission case, the outer peripheral portion of the cylindrical portion and the transmission case A plurality of second friction engagement elements disposed between the inner peripheral portions are provided. Since at least a part of the first friction engagement element and the second friction engagement element sharing the cylindrical portion of the clutch drum overlap in the axial direction, the total axial dimension and radial direction of the brake and the clutch are reduced. At the same time, the weight can be reduced. Since the ring gear and the cylindrical portion extend in the opposite directions along the axial direction from the coupling portion that couples them, the load from the ring gear is radially inward from the cylindrical portion via the first friction engagement element and the clutch hub. The transmission of the load is greatly simplified, and the radial dimension and axial dimension of the clutch can be reduced and the weight can be reduced. In addition, since there is no brake on the outside of the ring gear in the radial direction, the outer diameter of the planetary gear mechanism can be increased to increase the ratio width, which can reduce engine fuel consumption and improve vehicle running performance. become.

  According to the second aspect of the present invention, the forward movement 9 is achieved by engaging at least three frictional engagement means among the seven frictional engagement means including the first to third clutches and the first to fourth brakes. It is possible to establish a shift stage that is greater than or equal to the stage.

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 thrust load F1, F2, F3, F4 in the 3rd speed and 4th speed gear stage. The figure which shows the comparative example corresponding to FIG.

  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 extends in the axial direction from the radially outer end of the clutch drum 39, the clutch drum 39 integrally connected to the fourth ring gear Rd, the clutch hub 40 integrally connected to the input shaft 12. A plurality of friction engagement elements 41... Disposed between the cylindrical portion 39 a and the clutch hub 40 and a clutch piston 42 slidably disposed inside the clutch drum 39 are provided and supplied to the oil chamber 43. When the clutch piston 42 is driven by hydraulic pressure to engage the friction engagement elements 41 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 is slidably disposed within the transmission case 13 and the plurality of friction engagement elements 44 disposed between the cylindrical portion 39a of the clutch drum 39 and 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. Is done.

  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. The fourth planetary gear mechanism PGSd, the third planetary gear mechanism PGSc, and the first planetary gear mechanism are sequentially arranged from top to bottom. 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.

  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. 4, and the tenth speed gear stage is established.

  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 with reference to 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.

  FIG. 16 shows a comparative example corresponding to FIG. 3, and the reference numerals of the members of the comparative example basically correspond to the reference numerals of the members of the embodiment. The comparative example includes an arrangement of a fourth brake B4 that couples the fourth ring gear Rd of the fourth planetary gear mechanism PGSd to the transmission case 13, and an arrangement of the third clutch C3 that couples the fourth ring gear Rd to the input shaft 12. However, this is different from the embodiment shown in FIG.

  That is, in the embodiment shown in FIG. 3, the friction engagement element of the fourth brake B4 is formed on the outer peripheral portion of the cylindrical portion 39a of the outer periphery of the clutch drum 39 integrally coupled with the fourth ring gear Rd of the fourth planetary gear mechanism PGSd. 44 ... and the outer periphery of the friction engagement elements 41 ... of the third clutch C3 is supported on the inner periphery of the cylindrical portion 39a of the clutch drum 39. The friction engagement elements 44 of the fourth brake B4 and the friction engagement elements 41 of the third clutch C3 overlap in the axial direction when viewed from the inner side or the outer side in the radial direction, and the fourth planetary gear mechanism. The fourth ring gear Rd of PGSd and the cylindrical portion 39a of the clutch drum 39 extend in opposite directions along the axial direction from a coupling portion that couples them together.

  Therefore, the cylindrical portion 39a of the clutch drum 39 integral with the fourth ring gear Rd of the fourth planetary gear mechanism PGSd is attached to the transmission case 13 via the friction engagement element 44 of the fourth brake B4 located on the radially outer side. The cylindrical portion 39a of the clutch drum 39 can be coupled to the input shaft 12 via the friction engagement elements 41 of the third clutch C3 and the clutch hub 40 located on the radially inner side.

  On the other hand, in the comparative example shown in FIG. 16, the fourth planetary gear mechanism PGSd can be directly coupled to the transmission case 13 through the friction engagement elements 44 of the fourth brake B4 at the outer periphery of the fourth ring gear Rd. The structure of the fourth brake B4 is simplified. However, the third clutch C3 includes a clutch hub 40 extending radially inward from the fourth ring gear Rd of the fourth planetary gear mechanism PGSd and a clutch drum 39 extending radially outward from the input shaft 12. A friction engagement element is provided between the outer peripheral portion of the cylindrical portion 40a that protrudes in the axial direction from the radially intermediate portion of 40 and the inner peripheral portion of the cylindrical portion 39a that protrudes in the axial direction from the radially outer end of the clutch drum 39. 41... Are arranged.

  As a result, the load from the fourth ring gear Rd of the fourth planetary gear mechanism PGSd located on the radially outer side is transmitted radially inward to the cylindrical portion 40a of the clutch hub 40, and then the frictional force is transmitted from the cylindrical portion 40a. It is transmitted radially outward to the cylindrical portion 39a of the clutch drum 39 through the coupling elements 41, and then transmitted radially inward from the cylindrical portion 39a to the input shaft 12 via the clutch drum 39. There is a problem in that the load transmission path becomes complicated and long, leading to an increase in size of the third clutch C3.

  Further, since the friction engagement elements 44 of the fourth brake B4 and the friction engagement elements 41 of the third clutch C3 are arranged so as to be shifted in the axial direction when viewed from the inside or outside in the radial direction, the fourth brake B4. In addition, there is a problem that the total axial dimension of the third clutch C3 is increased. Moreover, since the friction engagement elements 44 of the fourth brake B4 are arranged on the outer side in the radial direction of the fourth ring gear Rd of the fourth planetary gear mechanism PGSd, the diameter of the fourth planetary gear mechanism PGSd is changed due to a change in gear ratio or the like. When the diameter of the frictional engagement elements 44 of the fourth brake B4 increases as a result of this, problems such as an increase in friction and an increase in the radial dimension of the automatic transmission T occur.

  On the other hand, according to the present embodiment shown in FIG. 3, the friction engagement elements 44 of the fourth brake B4 and the friction engagement elements 41 of the third clutch C3 are viewed from the inside or outside in the radial direction. Sometimes they overlap in the axial direction, and the frictional engagement elements 44 of the fourth brake B4 and the frictional engagement elements 41 of the third clutch C3 share the cylindrical portion 39a of the clutch drum 39 of the third clutch C3. Therefore, the total axial dimension and radial dimension of the fourth brake B4 and the third clutch C3 can be reduced, and the weight can be reduced. Further, since the friction engagement elements 44 of the fourth brake B4 are not arranged on the outer side in the radial direction of the fourth ring gear Rd of the fourth planetary gear mechanism PGSd, even if the diameter of the fourth planetary gear mechanism PGSd increases, the influence is increased. It does not reach the fourth brake B4.

  In addition, the load from the fourth ring gear Rd of the fourth planetary gear mechanism PGSd located on the radially outer side is applied to the friction engagement of the third clutch C3 from the cylindrical portion 39a of the clutch drum 39 in which the fourth ring gear Rd is extended in the axial direction. Since it is transmitted to the input shaft 12 radially inward via the elements 41... And the clutch hub 40, the load transmission path is greatly simplified, and the third clutch C3 can be reduced in size and weight.

  Furthermore, since the fourth brake B4 does not exist on the outer side in the radial direction of the fourth ring gear Rd of the fourth planetary gear mechanism PGSd, the outer diameter of the fourth planetary gear mechanism PGSd can be enlarged to increase the ratio width. As a result, the fuel consumption of the engine can be reduced and the running performance of the vehicle can be improved.

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

PSGa 1st planetary gear mechanism (planetary gear mechanism)
PSGb Second planetary gear mechanism (planetary gear mechanism)
PSGc 3rd planetary gear mechanism (planetary gear mechanism)
PSGd 4th planetary gear mechanism (planetary gear mechanism)
B1 First brake (brake)
B2 Second brake (brake)
B3 Third brake (brake)
B4 Fourth brake (brake)
C1 1st clutch (clutch)
C2 Second clutch (clutch)
C3 3rd clutch (clutch)
Ca first carrier (second element)
Cb second carrier (fifth element)
Cc 3rd carrier (8th element)
Cd 4th carrier (11th element)
Ra 1st ring gear (3rd element)
Rb 2nd ring gear (4th element)
Rc 3rd ring gear (9th element)
Rd 4th ring gear (10th element)
Sa First sun gear (first element)
Sb 2nd sun gear (6th element)
Sc 3rd sun gear (7th element)
Sd 4th sun gear (12th element)
12 Input shaft 13 Transmission case 29 Output gear (output member)
39 Clutch drum 39a Cylindrical portion 40 Clutch hub 41 Friction engagement element (first friction engagement element)
44 Friction engagement element (second friction engagement element)

Claims (2)

A plurality of planetary gear mechanisms having elements of a sun gear, a carrier, and a ring gear; and a plurality of engaging means for coupling the elements of the plurality of planetary gear mechanisms to each other or to other members. In an automatic transmission that establishes a plurality of shift stages according to engagement and disengagement of engagement means,
A multi-plate clutch (C3) for coupling one ring gear (Rd) of the plurality of planetary gear mechanisms to the input shaft (12), and a multi-plate type for coupling the ring gear (Rd) to the transmission case (13). Brake (B4)
The clutch (C3) includes a clutch hub (40) coupled to the input shaft (12), a clutch drum (39) having a cylindrical portion (39a) coupled to the ring gear (Rd), and the clutch hub. A plurality of first friction engagement elements (41) disposed between the outer periphery of (40) and the inner periphery of the cylindrical portion (39a),
The brake (B4) includes a plurality of second friction engagement elements (44) disposed between an outer peripheral portion of the cylindrical portion (39a) and an inner peripheral portion of the transmission case (13).
The first friction engagement element (41) and the second friction engagement element (44) at least partially overlap in the axial direction, and the clutch hub (40) includes the plurality of first friction engagement elements ( 41) is coupled to the input shaft (12) within the axial width of the shaft 41, and the ring gear (Rd) and the cylindrical portion (39a) are opposite to each other along the axial direction from the coupling portion coupling them. An automatic transmission characterized by extending. First to fourth planetary gear mechanisms (PGS1, PGS2, PGS3, PGS4), first to third clutches (C1, C2, C3), first to fourth brakes (B1, B2, B3, B4), With
The three elements of the first planetary gear mechanism (PGSa) are arranged in order at intervals corresponding to the gear ratio in the schematic diagram, and the first element (Sa), second element (Ca) and third element ( Ra)
The three elements of the second planetary gear mechanism (PGSb) are arranged at intervals corresponding to the gear ratio in the schematic diagram in order from the fourth element (Rb), the fifth element (Cb) and the sixth element ( Sb)
The three elements of the third planetary gear mechanism (PGSc) are arranged in order at intervals corresponding to the gear ratio in the schematic diagram, and the seventh element (Sc), eighth element (Cc) and ninth element ( Rc)
The tenth element (Rd), the eleventh element (Cd) and the twelfth element (from the one side) are arranged in order from the three elements of the fourth planetary gear mechanism (PGSd) at intervals corresponding to the gear ratio in the schematic diagram. Sd)
The seventh element (Sc) is connected to the input shaft (12), the fourth element (Rb) is connected to an output member (29), and the eighth element (Cc) and the eleventh element (Cd) And the third element (Ra) are connected to form a first connected body, the ninth element (Rc) and the sixth element (Sb) are connected to form a second connected body, The third element is configured by connecting the second element (Ca) and the fifth element (Cb),
The first clutch (C1) is capable of coupling the seventh element (Sc) and the third connector,
The second clutch (C2) can couple the twelfth element (Sd) and the second connector.
The third clutch (C3) can couple the seventh element (Sc) and the tenth element (Rd);
The first brake (B1) can couple the first element (Sa) to the transmission case (13),
The second brake (B2) can couple the third connecting body to the transmission case (13),
The third brake (B3) can couple the twelfth element (Sd) to the transmission case (13),
The fourth brake (B4) can couple the tenth element (Rd) to the transmission case (13),
The automatic transmission according to claim 1, wherein the ring gear is the tenth element (Rd).

Images