JP4765873B2 - Automatic transmission - Google Patents

Description

  The present invention relates to a structure of an automatic changer that engages and disengages a multi-plate clutch with a hydraulic servo, and more particularly to an improvement in control stability of the hydraulic servo.

  Conventionally, an input shaft, a normally-reducing planetary gear set that constantly decelerates and outputs the rotation of the input shaft, a subsequent planetary gear set that receives the output rotation of the always-reducing planetary gear set, and an output unit of the always-reducing planetary gear set; An inner peripheral side of the multi-plate clutch, comprising: a multi-plate clutch provided between an input portion of the subsequent planetary gear set; a clutch piston that presses the multi-plate clutch; and a hydraulic servo that controls the clutch piston. An automatic transmission in which a hub portion provided on the outer peripheral side of the multi-plate clutch is connected to the input portion of the succeeding planetary gear set, while a hub portion provided on the outer periphery of the multi-plate clutch is connected to the output portion of the constant speed reduction planetary gear set Is known (see, for example, Patent Document 1).

In the automatic transmission shown in Patent Document 1, a clutch piston and a hydraulic servo (a general term for mechanisms such as a control hydraulic chamber and a return spring for operating the clutch piston) are provided on the drum side.
JP 2002-349647 A

  However, when the hydraulic servo is provided on the drum side in the automatic transmission having the above-described conventional structure, the automatic transmission may be easily affected by fluctuations in the rotation of the drum. This is because the rotation of the drum is the input rotation of the subsequent planetary gear set, and the input rotation speed (speed ratio with respect to the input shaft) often varies greatly depending on the gear position. For example, in Patent Document 1, a clutch piston for a multi-plate clutch (clutch C3 in the same document) and a hydraulic servo are provided on the drum side, but the drum is stopped by the brake B1 in the same document at the second speed. Yes. Then, at the third speed, the clutch C3 is engaged and the speed is increased to a predetermined rotational speed (the output rotational speed of the constant reduction planetary gear set). In other words, even if the input rotation speed is constant, the rotation speed of the clutch piston and the hydraulic servo greatly changes due to the shift.

  Normally, when the hydraulic servo rotates, centrifugal hydraulic pressure is generated in the control hydraulic chamber. This centrifugal hydraulic pressure causes the control hydraulic pressure to vary during or during clutch engagement. Further, when the clutch is not engaged, the clutch piston malfunctions and the outer peripheral piston seal blows out. In any case, the control stability of the hydraulic servo is lowered, which is not preferable.

  As a mechanism for reducing the influence of centrifugal hydraulic pressure, for example, a balanced piston mechanism shown in Patent Document 1 is well known. The balance piston mechanism is a mechanism in which a hydraulic chamber (balance hydraulic chamber) is provided on the side where the clutch piston is opened, and hydraulic pressure (such as lubricating oil) having a source pressure as small as possible is guided to the balance hydraulic chamber. According to this mechanism, since the centrifugal hydraulic pressure in the control hydraulic chamber and the centrifugal hydraulic pressure in the balance hydraulic chamber are offset, it is possible to obtain an effect as if the centrifugal hydraulic pressure apparently disappears or becomes small.

  However, it is difficult to completely eliminate the influence of the centrifugal hydraulic pressure even with the balance piston mechanism. Especially in the case of a transient change with a large rotational fluctuation as described above, the influence of the centrifugal hydraulic pressure cannot always be reduced sufficiently. . Further, even if the centrifugal force oil pressure is apparently reduced, the centrifugal force oil pressure is not actually reduced, so it cannot be an effective measure against the blowout of the outer peripheral side piston seal, for example.

  The present invention has been made in view of the above circumstances, and provides an automatic transmission capable of effectively suppressing the influence of centrifugal hydraulic pressure acting on a hydraulic servo and thereby improving the control stability of the hydraulic servo. The purpose is to do.

The invention according to claim 1 for solving the above-mentioned problem is an input shaft, a constantly-decelerating planetary gear set that constantly decelerates and outputs the rotation of the input shaft, and a subsequent input that receives the output rotation of the always-reducing planetary gear set A planetary gear set, a multi-plate clutch provided between an output portion of the constantly reducing planetary gear set and an input portion of the subsequent planetary gear set, a clutch piston for pressing the multi-plate clutch, and a hydraulic pressure for controlling the clutch piston A servo and a boss portion that is fixed to the transmission case and extends in the input shaft direction and through which the input shaft is inserted , and a hub portion provided on an inner peripheral side of the multi-plate clutch is configured to be the constant reduction planetary gear The drum connected to the output portion of the set and provided on the outer peripheral side of the multi-plate clutch is connected to the subsequent planetary gear. An automatic transmission that is coupled to the input of the set, the clutch piston and the hydraulic servo is provided on the member connected to the hub portion or the hub portion, the above constant deceleration planetary gear set ring gear An input and a carrier output, wherein a sun gear is fixed to the boss portion, and the hub portion is in communication with the carrier and a sleeve provided rotatably on the boss portion, and The clutch piston is in sliding contact with the inner peripheral surface of the hub portion and the outer peripheral surface of the sleeve .

  According to a second aspect of the present invention, in the automatic transmission according to the first aspect, the hub portion is notched at a plurality of locations in the circumferential direction of a portion facing the piston, and a hub side fitting portion that is a notch remaining portion is provided. The clutch piston is formed with a piston-side notch portion in which a portion corresponding to the hub-side fitting portion is notched, and the hub portion and the clutch piston include the hub-side fitting portion and the piston. It is characterized by being arranged in an overlapping manner in the axial direction so that the side notch is fitted.

According to a third aspect of the present invention, in the automatic transmission according to the first or second aspect, the clutch piston is formed with a cylindrical portion having an inner peripheral surface coaxial with the sleeve outer peripheral surface and having a larger diameter than the sleeve outer peripheral surface. The clutch piston is coaxially adjacent, the inner peripheral portion is locked to the sleeve outer peripheral surface so as not to move away from the clutch piston, and the outer peripheral portion is the above-described cylindrical portion of the clutch piston. A seal plate that is in sliding contact with an inner peripheral surface, wherein a control hydraulic chamber of the hydraulic servo is configured such that the seal plate and the clutch piston face each other, the inner peripheral surface of the cylindrical portion of the clutch piston, and the sleeve It is characterized by being formed with the outer peripheral surface.

According to a fourth aspect of the present invention, in the automatic transmission according to the third aspect , the balance hydraulic chamber of the hydraulic servo is configured so that the clutch piston and the carrier face each other, the inner peripheral surface of the hub portion, and the It is formed with the outer peripheral surface of a sleeve.

According to a fifth aspect of the present invention, in the automatic transmission according to the fourth aspect , the carrier supports a pinion shaft that supports a pinion of the constant speed planetary gear set, and the pinion shaft lubricates the pinion. A pinion lubricating oil passage is formed, and a portion forming the balance hydraulic chamber of the carrier communicates with the pinion lubricating oil passage.

According to a sixth aspect of the present invention, in the automatic transmission according to any one of the first to fifth aspects, the multi-portion is provided between the outer peripheral surface of the hub portion and the inner peripheral surface with which the clutch piston slides. A clutch lubricating oil passage extending in the direction in which the plate clutches are arranged is formed, and a portion where the balance hydraulic chamber is not formed on the inner peripheral surface of the hub portion and the clutch lubricating oil passage are connected to each other. A first lubrication hole that communicates, and a second lubrication hole that communicates the clutch lubricant oil passage and the outer peripheral surface of the hub portion are provided.

  According to the first aspect of the present invention, as described below, the influence of the centrifugal hydraulic pressure acting on the hydraulic servo can be effectively suppressed, and the control stability of the hydraulic servo can be improved.

  According to the present invention, since the clutch piston and the hydraulic servo are provided on the hub portion or a member connected to the hub portion, the clutch piston and the hydraulic servo rotate integrally with the hub portion. Furthermore, since this hub part is connected to the output part of the always-reducing planetary gear set, the clutch piston and the hydraulic servo eventually rotate integrally with the output part of the always-reducing planetary gear set.

  The output part of the constantly reducing planetary gear set is a stable rotating element that always rotates at a constant rate with respect to the input shaft. That is, it does not stop at a shift stage, becomes faster than the input shaft, and does not frequently repeat such rotation fluctuations. There is no rotational fluctuation even during gear shifting.

  Therefore, the hydraulic servo according to the present invention hardly undergoes a transient influence because there is almost no fluctuation in the centrifugal hydraulic pressure associated with the shift. Further, since the rotation speed of the hydraulic servo is always smaller than the input shaft rotation speed, the centrifugal force hydraulic pressure itself is always small, and it is difficult to receive adverse effects (for example, blowout of the outer peripheral side piston seal) due to excessive centrifugal force hydraulic pressure. Thus, according to the present invention, the control stability of the hydraulic servo is effectively enhanced.

  Note that the present invention is not intended to exclude other centrifugal force hydraulic pressure reduction measures such as the balance piston mechanism, and can be used in combination with them, and when used in combination with them, the effect can be enhanced synergistically. Is.

  As an additional effect of the present invention, the automatic transmission can be made compact with a hydraulic servo at the center. Since the clutch piston of the present invention is provided on the hub portion side as described above, the conventional general structure, that is, the structure in which the clutch piston is stored in the piston drum and the clutch piston is operated therein (patent) (Ref. 1). That is, the piston drum can be omitted as a component requirement of the hydraulic servo, and the size can be reduced accordingly.

Furthermore, according to the present invention, by arranging the sliding contact portion of the clutch piston on the inner peripheral side of the hub portion, it is possible to further shorten the axial direction and make it more compact.

  According to the invention of claim 2, as described below, it is possible to further promote downsizing around the hydraulic servo.

  When the clutch piston is provided on the drum side as in the conventional structure, relative rotation occurs between the clutch piston and the hub portion. That is, an axial clearance (clearance) for preventing interference is required between the clutch piston and the hub portion. This clearance is the minimum when the clutch piston is fully stroked to the hub side, but this minimum clearance must be ensured to be 0 or more even when the variation is swung to the worst side. I must.

  However, according to the present invention, since the clutch piston is provided on the hub portion side and there is no relative rotation, there is no need to provide such a clearance. Therefore, the minimum clearance can be easily reduced to 0 by arranging the hub portion and the clutch piston so as to overlap in the axial direction. That is, the axial direction can be shortened and made compact accordingly.

According to the invention of claim 3 , by arranging the control hydraulic chamber of the hydraulic servo on the inner peripheral side of the hub portion, it is possible to further shorten the axial direction and make it more compact.

According to the fourth aspect of the present invention, a balance piston mechanism having a balance hydraulic chamber can be taken, and the influence can be further reduced by reducing the apparent centrifugal force hydraulic pressure. That is, the control stability of the hydraulic servo can be further improved.

  Further, by disposing the balance hydraulic chamber on the inner peripheral side of the hub portion, it is possible to further shorten the axial direction and make it more compact.

  Further, the balance hydraulic chamber is configured by utilizing a part of the carrier (surface facing the clutch piston). That is, since a separate seal plate (see Patent Document 1) for forming a balanced hydraulic chamber as seen in the conventional general structure can be omitted, the structure can be simplified and made compact. .

According to the invention of claim 5 , by utilizing the fact that a part of the carrier constitutes the balance hydraulic chamber, the lubricating oil to the pinion is easily guided by a simple structure that passes through the balance hydraulic chamber and the pinion lubricating oil passage. be able to.

According to the invention of claim 6 , as described below, the multi-plate clutch can be properly lubricated.

  Usually, lubrication of the multi-plate clutch is performed by introducing lubricating oil into a lubricating hole penetrating from the inner peripheral side to the outer peripheral side of the hub portion. However, when the clutch piston is slidably brought into contact with the inner peripheral surface of the hub portion as in the present invention, a lubrication hole cannot be provided in that portion, and there is a risk that the lubrication of the multi-plate clutch will be insufficient.

  Therefore, according to the present invention, a lubricating oil passage extending in the direction in which the multi-plate clutch is arranged is formed, and the first lubricating hole and the second lubricating hole are provided, so that no balance hydraulic chamber is formed on the inner peripheral surface of the hub portion. Lubricating oil can be guided from the portion over the entire axial direction of the multi-plate clutch via the first lubricating hole, the clutch lubricating oil passage, and the second lubricating hole.

  Moreover, in the case of such a structure, the second lubricating hole does not open on the inner peripheral surface of the hub portion, so that the operation of the clutch piston is not affected.

  From the viewpoint of reducing the radial direction, it is desirable that the radial length of the clutch lubricating oil passage be as short as possible. Therefore, for example, the part including the outer peripheral part of the hub part is a ring-shaped member separate from the part including the inner peripheral part, and both of them are fitted so that a slight gap is formed between the inner and outer peripherals of both, It is also preferable that the gap be a clutch lubricating oil passage.

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

  FIG. 1 is a schematic configuration diagram of an automatic transmission according to an embodiment of the present invention. The automatic transmission 1 is applied to a horizontal engine, and has, as main components, a torque converter 10 driven by the engine and a transmission mechanism 20 to which an output rotation of the torque converter 10 is input. The output rotation of the transmission mechanism 20 is input to the differential device 52 via the intermediate transmission mechanism 50.

  The torque converter 10 includes a case 11 connected to the engine output shaft 2, a pump 12 fixed in the case 11, a pump 12 disposed opposite to the pump 12, and driven by hydraulic fluid through the pump 12. A stator 15 that is interposed between the pump 12 and the turbine 13 and that is supported by the transmission case 3 via the one-way clutch 14 to increase torque, and the case 11 and the turbine 13. The lockup clutch 16 is provided between the engine output shaft 2 and the turbine 13 via the case 11. The rotation of the turbine 13 is output to the transmission mechanism 20 side via the turbine shaft 17 (input shaft).

  An oil pump 18 that is driven by the engine output shaft 2 via the case 11 of the torque converter 10 is disposed on the opposite side of the torque converter 10 from the engine.

  The speed change mechanism 20 includes a first planetary gear set 21 disposed on the torque converter 10 side, a second planetary gear set 22 disposed on the anti-torque converter 10 side, and third and fourth planetary gears disposed between them. And a plurality of friction elements such as clutches and brakes for switching the power transmission path including each of the planetary gear sets 21 to 24. The first and first are sequentially arranged from the torque converter 10 side. 2, third clutches 31, 32, 33, and first and second brakes 41, 42, which are configured to obtain forward 1 to 6 speeds and reverse speeds by selective interruption. .

  The first to fourth planetary gear sets 21 to 24 all have sun gears 21a to 24a, a plurality of pinions 21b to 24b that mesh with these sun gears 21a to 24a, and carriers 21c to 21c that support these pinions 21b to 24b. 24c and ring gears 21d to 24d meshing with the pinions 21b to 24b.

  In the first planetary gear set 21, the sun gear 21a is fixed to the transmission case 3 (or a member fixed thereto), the carrier 21c is connected to the hub portion of the first clutch 31, and the ring gear 21d is connected to the turbine shaft 17 and the first gear. The two clutches 32 are connected to the hub portion.

  In the second planetary gear set 22, the sun gear 22 a is fixed to the transmission case 3 (or a member fixed thereto), the carrier 22 c is connected to the hub portion of the third clutch 33, and the ring gear 22 d is connected to the turbine shaft 17. It is connected.

  In the third planetary gear set 23, the sun gear 23 a is connected to the clutch drum of the third clutch 33 and the hub portion of the first brake 42, and the carrier 23 c is connected to the clutch drum of the second clutch 32 and the hub portion of the first brake 41. The ring gear 23d is connected to the carrier 24c of the fourth planetary gear set 24.

  Further, in the fourth planetary gear set 24, the sun gear 24a is connected to the clutch drum of the first clutch 31, and the ring gear 24d is connected to the carrier 23c of the third planetary gear set 23 to be connected to the hub portion of the first brake 41. 24 c is coupled to the ring gear 23 d of the third planetary gear set 23 and coupled to the output gear 25 to the intermediate transmission mechanism 50.

  The rotation of the output gear 25 is input to the differential device 52 via the intermediate transmission mechanism 50 and drives the left and right axles 55 and 56.

  The automatic transmission 1 having the schematic configuration as described above is shown in Table 1 according to combinations of on (engaged) and off (released) of the first to third clutches 31 to 33 and the first and second brakes 41 and 42. Thus, forward 1st to 6th speed and reverse speed are realized.

  In Table 1, ◯ indicates that each clutch or brake is in an engaged state, and no symbol indicates that it is in a released state. That is, the first clutch 31 and the second brake 42 are engaged at the first speed, the first clutch 31 and the first brake 41 are engaged at the second speed, and the first clutch 31 and the third clutch 33 are engaged at the third speed, In the fourth speed, the first clutch 31 and the second clutch 32 are engaged. In the fifth speed, the second clutch 32 and the third clutch 33 are engaged. In the sixth speed, the second clutch 32 and the first brake 41 are engaged. Then, the third clutch 33 and the second brake 42 are engaged.

  Each clutch and brake is turned on and off by supplying and discharging hydraulic pressure to the corresponding hydraulic servo. The hydraulic pressure is supplied and discharged by a control valve (not shown). The control valve is an assembly of a plurality of spool valves such as a pressure regulating valve for setting the hydraulic pressure and a shift valve for switching the supply and discharge of the hydraulic pressure to the hydraulic servo. In this embodiment, the spool valve is controlled by a plurality of solenoid valves. A control unit 90 for electrically controlling the solenoid valve is provided.

  The control unit 90 is a control unit that controls the operation of the automatic transmission 1 according to the operating state of the engine, the vehicle, and the automatic transmission 1 itself, and includes a computer having a CPU, a ROM, a RAM, and the like. Specifically, the operation of each solenoid valve is controlled by a CPU executing a program stored in advance in a ROM (or RAM).

  The control unit 90 receives detection signals from various sensors such as an engine rotation speed sensor and an oil temperature sensor that detects the temperature of hydraulic oil. FIG. 1 shows a first rotation speed sensor 92, a second rotation speed sensor 94, and a third rotation speed sensor 96 among the various sensors. These rotational speed sensors are sensors that detect rotational speeds of various parts of the speed change mechanism 20, and the control unit 90 can execute precise shift control by detecting the rotational speeds.

  FIG. 2 is a partial cross-sectional view (upper half) of the automatic transmission 1 and shows the vicinity of the first clutch 31, the second clutch 32, and the output gear 25.

  The oil pump 18 is a member fixed to the transmission case 3, and FIG. 2 shows an oil pump cover 18 a constituting the oil pump 18. A part of the oil pump cover 18a extends in the axial direction to form a cylindrical boss portion 18b through which the turbine shaft 17 is inserted. That is, the boss portion 18b is fixed to the transmission case 3 via the oil pump cover 18a (oil pump 18).

  The sun gear 21a of the first planetary gear set 21 is fixed to the boss portion 18b. On the other hand, a ring gear 21 d is connected to the turbine shaft 17. Therefore, when the turbine shaft 17 rotates, the ring gear 21d rotates around the fixed sun gear 21a at the same rotation. The pinion 21b rolls around the fixed sun gear 21a while rotating about its own axis. Accordingly, the carrier 21c that supports the pinion 21b rotates around the sun gear 21a at a constant rate slower than the rotational speed of the ring gear 21d, that is, the rotational speed of the turbine shaft 17. Since the output element of the first planetary gear set 21 is the carrier 21c, the first planetary gear set 21 is a constant reduction planetary gear set that always decelerates and outputs the rotation of the input shaft (turbine shaft 17).

  A feature of the rotation of the carrier 21c that is always decelerated is that the carrier 21c always rotates at a constant ratio with respect to the turbine shaft 17. It does not stop or rotate according to the gear position, and does not change faster or slower than the turbine shaft 17. In other words, the rotation is stable without sudden fluctuation.

  The second feature is that the rotation is always relatively low. This is naturally a planetary gear set that decelerates the rotation of the turbine shaft 17 at all times. However, the fact that the rotation is always lower than the turbine shaft 17 has the advantage that it is less susceptible to centrifugal force due to high rotation. Have.

  As shown in FIG. 1, the first clutch 31 includes a multi-plate clutch provided between a carrier 21 c that is an output portion of the first planetary gear set 21 and a sun gear 24 a of the fourth planetary gear set 24 that follows the carrier 21 c. . When the first clutch 31 is engaged, the output portion (carrier 21c) of the first planetary gear set 21 and the input portion (sun gear 24a) of the fourth planetary gear set 24 are coupled. That is, the fourth planetary gear set 24 is a subsequent planetary gear set that receives the output rotation of the first planetary gear set 21 as an input.

  As shown in FIG. 2, in the first clutch 31, a hub portion 80 provided on the inner peripheral side of the multi-plate clutch and the carrier 21c are connected. On the other hand, a drum 89 provided on the outer peripheral side of the multi-plate clutch and a sun gear 24a of a fourth planetary gear set 24 (not shown) are connected.

  The first clutch 31 includes a clutch piston 61 that presses the multi-plate clutch and a hydraulic servo 60 that controls the clutch piston 61. The clutch piston 61 and the hydraulic servo 60 are provided on the hub portion 80 or a member (carrier 21c or sleeve 28) connected to the hub portion 80.

  A sensing rotor 77 extending integrally with the clutch piston 61 is provided on the outer peripheral side of the clutch piston 61 (see FIG. 5). The sensing rotor 77 is a detected part of the first rotation speed sensor 92. The sensing rotor 77 has a cylindrical portion that surrounds the drum 89 with a slight gap, and a notch 77a is formed in the cylindrical portion. A large number of notches 77 a are formed at equal intervals in the circumferential direction of the cylindrical portion of the sensing rotor 77.

  As shown in FIG. 2, a first rotation speed sensor 92 is provided at a position corresponding to the outer peripheral side of the drum 89 and the notch 77 a of the sensing rotor 77. The first rotation speed sensor 92 detects that the notch 77a has passed through the detection area. The control unit 90 detects the rotational speed of the sensing rotor 77 and, consequently, the rotational speed of the carrier 21c (the output rotational speed of the first planetary gear set 21) by counting the number of passages of the notch 77a per unit time. Further, as described above, since the rotation of the carrier 21c is a reduction in the rotation of the turbine shaft 17 at a constant rate, the control unit 90 detects the rotation speed (input rotation speed) of the turbine shaft 17 by a simple calculation. can do.

  Since the control unit 90 detects the rotational speed of the turbine shaft 17 based on the rotational speed of the carrier 21c that rotates stably, the controllability of the hydraulic servo 60 can be improved.

  A second rotation speed sensor 94 is provided adjacent to the first rotation speed sensor 92. Equally spaced holes or irregularities are provided at equal intervals in the circumferential direction at positions where the drum 89 faces the tip of the second rotation speed sensor 94. The second rotation speed sensor 94 detects that the unevenness has passed through the detection area. The control unit 90 can detect the rotational speed of the drum 89, and hence the rotational speed of the sun gear 24a of the fourth planetary gear set 24, by counting the number of passages per unit time. Thus, by detecting the rotational speed (intermediate part rotational speed) of the member between the input part (turbine shaft 17) and the output part (output gear 25), more accurate shift control can be performed. Furthermore, the controllability of the hydraulic servo 60 can be improved.

  A third rotation speed sensor 96 is provided close to or adjacent to the first rotation speed sensor 92 and the second rotation speed sensor 94. The tip of the third rotational speed sensor 96 faces the output gear 25, and the third rotational speed sensor 96 detects that one gear tooth has passed through the detection area using the unevenness of the gear. To do. The control unit 90 can detect the rotational speed of the output gear 25 (the output rotational speed of the fourth planetary gear set 24) by counting the number of passing teeth per unit time. As described above, by detecting the rotation speed of the output unit, it is possible to perform shift control with higher accuracy, and to further improve the controllability of the hydraulic servo 60.

  Further, the first rotation speed sensor 92, the second rotation speed sensor 94, and the third rotation speed sensor 96 are modularized (integrated) by utilizing the fact that they are disposed close to each other, and the control unit 90 is integrated. Built in. By doing so, it is possible to improve these assembling properties and increase the productivity of the automatic transmission 1. Further, since each of these sensors is directly incorporated in the control unit 90, it is less likely to pick up noise than in the case where the harness is coupled at a distant position, and more accurate detection is possible.

  FIG. 3 is a partially enlarged view of FIG. 2 and shows the periphery of the first planetary gear set 21 and the first clutch 31 in more detail.

  On the inner peripheral side of the carrier 21c, a ring-shaped sleeve 28 that fits into the boss portion 18b is formed. The inner ring of the bearing 19 is fixed to the boss portion 18b, and the outer ring of the bearing 19 is fixed to the sleeve 28. Therefore, the sleeve 28 is freely rotatable around the boss portion 18b.

  An extended ring 28 a is integrally fitted on the outer peripheral side of the sleeve 28. As a result, the outer peripheral surface of the sleeve 28 including the extended ring 28a has a two-stage structure of a small-diameter side outer peripheral surface and a large-diameter side outer peripheral surface (the outer peripheral surface of the extended ring 28a).

  A clutch piston 61 is provided on the outer peripheral surface of the large diameter side of the sleeve 28 so as to be in sliding contact with the seal 28 via a seal member. The clutch piston 61 is a bent disk-like member, and a cylindrical portion 61b having an inner peripheral surface coaxial with the outer peripheral surface of the sleeve 28 is formed slightly closer to the outer periphery than the middle in the radial direction.

  A substantially disc-shaped seal plate 71 is provided on the side of the clutch piston 61 on the axial oil pump 18 side. The inner peripheral side of the seal plate 71 is fitted to the outer peripheral surface on the small diameter side of the sleeve 28 via a seal member, and is locked by a snap ring 75 so as not to move away from the clutch piston 61. The outer peripheral side of the seal plate 71 is in sliding contact with the inner peripheral surface of the cylindrical portion 61b of the clutch piston 61 via a seal member. Accordingly, the control hydraulic chamber 62 of the hydraulic servo 60 includes the surfaces of the seal plate 71 and the clutch piston 61 facing each other, the inner peripheral surface of the cylindrical portion 61b of the clutch piston 61, and the outer peripheral surface (small-diameter side outer peripheral surface) of the sleeve 28. And is formed.

  When hydraulic pressure is supplied from a control valve (not shown) to the control hydraulic chamber 62 via an oil passage formed in the oil pump cover 18a and the sleeve 28, the clutch piston 61 is turned on (direction away from the seal plate 71). Stroke to. A plurality of concentric return springs 65 (coil springs) are provided between the clutch piston 61 and the carrier 21c. The return spring 65 always urges the clutch piston 61 to the off side, and prevents the clutch piston 61 from moving to the on side when the clutch is turned off.

  A hub portion 80 is formed on the outer peripheral side of the carrier 21c. The hub portion 80 is a member that is provided on the inner peripheral side of the friction plate 82 constituting the multi-plate clutch and supports the friction plate 82 while guiding it. The hub portion 80 is a substantially annular member that is coaxial with the sleeve 28 as a whole. A cylindrical portion 86 is formed on the inner peripheral side of the hub portion 80 of the carrier 21 c and in the vicinity of the clutch piston 61. The outer peripheral surface of the cylindrical portion 61b of the clutch piston 61 is in sliding contact with the inner peripheral surface of the cylindrical portion 86 through the seal member.

  Thus, since the clutch piston 61 and the hydraulic servo 60 are provided on the hub portion 80 or a member connected to the hub portion 80, the clutch piston 61 and the hydraulic servo 60 rotate integrally with the hub portion 80. That is, the clutch piston 61 and the hydraulic servo 60 rotate integrally with the carrier 21c.

  As described above, the carrier 21c is a stable rotating element that always rotates at a constant rate relative to the turbine shaft 17. Therefore, the hydraulic servo 60 hardly undergoes a transient influence because there is almost no fluctuation in the centrifugal hydraulic pressure associated with the shift. Further, since the rotational speed of the hydraulic servo 60 is always smaller than the rotational speed of the turbine shaft 17, the centrifugal force hydraulic pressure itself is always small, and it is difficult to receive adverse effects (for example, blowout of the outer peripheral piston seal) due to excessive centrifugal force hydraulic pressure. Thus, in this embodiment, the control stability of the hydraulic servo 60 is effectively enhanced.

  Furthermore, since the clutch piston 61 of this embodiment is provided on the hub portion 80 side as described above, the clutch piston is stored in the conventional general structure, that is, the piston drum, and the clutch piston is operated therein. It is not necessary to take the structure. That is, the piston drum is omitted as a component of the hydraulic servo 60, and the size is reduced accordingly.

  Further, by arranging the sliding contact portion of the clutch piston 61 and the control hydraulic chamber 62 of the hydraulic servo 60 on the inner peripheral side of the hub portion 80, the axial direction can be further shortened and made compact.

  Further, the hydraulic servo 60 balances between the mutually facing surfaces of the clutch piston 61 and the carrier 21c, the inner peripheral surface of the hub portion 80 (hub portion inner peripheral side member 86), and the outer peripheral surface of the sleeve 28 (large diameter outer peripheral surface). A hydraulic chamber 67 is formed. The balance hydraulic chamber 67 is supplied with very low pressure lubricating oil.

  Thus, the apparent centrifugal force hydraulic pressure acting on the hydraulic servo 60 is reduced by the balance piston mechanism including the balance hydraulic chamber 67. That is, since the centrifugal hydraulic pressure substantially equal to the centrifugal hydraulic pressure acting on the control hydraulic pressure chamber 62 across the clutch piston 61 also acts on the balance hydraulic pressure chamber 67, both balance (cancel), and the centrifugal hydraulic pressure is reduced. Such an effect is obtained. That is, the control stability of the hydraulic servo can be further improved.

  Further, by arranging the balance hydraulic chamber 67 on the inner peripheral side of the hub portion 80, the axial direction can be shortened and made compact.

  Further, the balance hydraulic chamber 67 is configured using a part of the carrier 21c (a surface facing the clutch piston 61). That is, since a separate seal plate for forming a balanced hydraulic chamber as in a conventional general structure is omitted, the structure is simplified and made compact.

  A pinion lubricating oil passage 27a for lubricating the pinion 21b is formed on the pinion shaft 27 that supports the pinion 21b, and a portion forming the balance hydraulic chamber 67 of the carrier 21c and the pinion lubricating oil passage 27a communicate with each other. ing.

  Thus, by utilizing the fact that a part of the carrier 21c constitutes the balance hydraulic chamber 67, the lubricating oil to the pinion 21b can be easily made with a simple structure that passes through the balance hydraulic chamber 67 and the pinion lubricating oil passage 27a. Guided.

  Next, a fitting structure between the hub portion 80 and the clutch piston 61 will be described.

  FIG. 4 is a perspective view of the hub portion 80 viewed from the oil pump 18 side (clutch piston 61 side). A hub part main body 81 which is a main body part of the hub part 80 is formed with a spline 81c extending in the axial direction. The hub part main body 81 is notched at a plurality of positions (four positions in the present embodiment) in the circumferential direction of the part facing the clutch piston 61 to form a hub side notch part 81a. And the remaining part (part not cut out) is the hub side fitting part 81b.

  FIG. 5 is a perspective view showing the clutch piston 61 and the hub portion 80 in the assembled state. As shown in this figure, a notch (through hole in the axial direction) is provided at a position corresponding to the hub side fitting portion 81b of the clutch piston 61, and a piston side notch 61a is formed. The hub portion 80 and the clutch piston 61 are overlapped in the axial direction so that the hub-side fitting portion 81b and the piston-side notch portion 61a are fitted. The clutch piston 61 strokes in the axial direction while maintaining such an overlapping state.

  Such an overlapping arrangement is possible because the clutch piston 61 is provided on the hub portion 80 side. That is, there is no relative rotation between the clutch piston 61 and the hub portion 80. If there is relative rotation on both sides, it is necessary to provide a clearance to avoid mutual interference, but this is not necessary in this embodiment. Therefore, the substantial clearance can be easily reduced to zero by arranging the hub portion 80 and the clutch piston 61 so as to overlap in the axial direction (minus clearance). That is, the axial direction is shortened and made compact accordingly.

  Further, with such a structure, the integral rotation of the hub portion 80 and the clutch piston 61 is further ensured, and the detection accuracy of the first rotation speed sensor 92 can be increased.

  As shown in FIG. 5, the clutch piston 61 is provided with a protrusion 61 c that protrudes to the inside of each piston-side notch 61 a and is fitted to the spline 81 c of the hub portion 80. By doing so, the play in the rotational direction between the hub portion 80 and the clutch piston 61 is packed as much as possible, so that the detection accuracy of the first rotational speed sensor 92 is further enhanced.

  Next, the multi-plate clutches 82 and 83 of the first clutch 31 will be described. The multi-plate clutches 82 and 83 are formed by alternately arranging the friction plates 82 and the disk plates 83 in the axial direction. Each of these plates has a slight gap (clutch clearance) in the axial direction in the released state. The friction plate 82 is obtained by attaching a friction material (facing material) to both surfaces of a metal annular plate. On the surface of the friction material, in order to reduce drag resistance when the clutch is released, a groove for accelerating the flow of the lubricating oil and cutting the oil film is formed. On the inner peripheral side of the friction plate 82, irregularities that fit into the splines 81 c of the hub portion 80 are formed, and rotate integrally with the hub portion 80. The friction plate 82 is movable along the spline 81c within the clutch clearance range.

  On the other hand, the disk plate 83 is a metal ring plate. On the outer peripheral side of the disk plate 83, irregularities that fit into splines 89 a formed on the inner peripheral side of the drum 89 are formed, and rotate integrally with the drum 89. The disc plate 83 is movable along the spline 89a within the clutch clearance range.

  However, the disk plates 83 a at both ends of the disk plate 83 are provided on the hub portion 80 side, like the friction plate 82. The end disk plate 83a is thicker than other disk plates 83 provided in the middle in order to prevent falling and heat absorption.

  The end disk plate 83a is affixed with a friction material only on one inner side. This is a contrivance for providing the friction plate 82 on the hub portion 80 side while the clutch piston 61 is provided on the hub portion 80 side. That is, when the clutch piston 61 is provided on the hub portion 80 side, the friction plate 82 is provided on the drum 89 side in a general arrangement (an arrangement in which no friction material is attached to the end disk plate 83a). Accordingly, the friction plate 82 can be provided on the hub portion 80 side. By doing so, the friction material of the friction plate 82 can be arranged on the side of the hub portion 80 to which the lubricating oil is guided and always stably rotates, so that the effect of the groove (reduction of drag resistance) is remarkably obtained. be able to.

  When the first clutch 31 is turned on by the hydraulic servo 60, the clutch piston 61 strokes toward the hub portion 80, closes the clutch clearance, and presses the multi-plate clutches 82 and 83. The pressed multi-plate clutches 82 and 83 are fastened to each other via a friction material, and the hub portion 80 and the drum 89 are fastened as a whole.

  Next, the lubrication mode of the multi-plate clutches 82 and 83 will be described. Usually, lubrication of the multi-plate clutch is performed by introducing lubricating oil into a lubricating hole penetrating from the inner peripheral side to the outer peripheral side of the hub portion. However, when the clutch piston 61 is brought into sliding contact with the inner peripheral surface of the hub portion 80 as in the present embodiment, a lubrication hole cannot be provided in that portion, and the multi-plate clutches 82 and 83 are insufficiently lubricated. There is a fear.

  Therefore, in this embodiment, the hub portion 80 has a two-layer structure. That is, the inner peripheral portion of the hub portion 80 is formed by the carrier 21c, and the outer peripheral portion is a ring-shaped member (hub portion main body 81) separate from the carrier 21c. The gap is used as a clutch lubricating oil passage 88.

  Then, the first lubricating hole 87 that communicates the portion where the balance hydraulic chamber 67 is not formed on the inner peripheral surface of the hub portion 80 and the clutch lubricating oil passage 88, and the clutch lubricating oil passage 88 and the outer peripheral surface of the hub portion 80 communicate. The second lubrication hole 85 is provided. By doing so, the multi-plate clutches 82 and 83 are passed through the first lubricating hole 87, the clutch lubricating oil passage 88 and the second lubricating hole 85 from the portion where the balance hydraulic chamber 67 is not formed on the inner peripheral surface of the hub portion 80. The lubricating oil is guided over the entire axial direction.

  Here, since the second lubricating hole does not open to the inner peripheral surface of the hub portion 80, the operation of the clutch piston 61 is not affected. In addition, since the hub portion 80 has a double structure, the radial length of the clutch lubricating oil passage 88 can be shortened, and the radial direction can be made compact. For example, it is advantageous in shortening the radial direction as compared with a method in which the clutch lubricating oil passage 88 is formed by drilling in the axial direction.

  As mentioned above, although embodiment of this invention was described, these embodiment can be suitably changed in the range which does not deviate from the summary of this invention. For example, the present invention is not limited to the automatic transmission having the schematic configuration shown in FIG. Further, the structure of the clutch piston 61 and the hydraulic servo 60 is not necessarily limited to the above structure, and any other structure is possible as long as it is provided on a hub portion or a member connected to the hub portion. .

1 is a schematic configuration diagram of an automatic transmission according to an embodiment of the present invention. It is a fragmentary sectional view of the automatic transmission. FIG. 3 is a partially enlarged view of FIG. 2. It is the perspective view which looked at the hub part of the above-mentioned automatic transmission from the oil pump. It is a perspective view which shows the clutch piston and hub part of the said automatic transmission in the assembly | attachment state.

1 Automatic transmission 3 Transmission case 17 Turbine shaft (input shaft)
18b Boss part 21 1st planetary gear set (always reducing planetary gear set)
21a Sun gear 21b Pinion 21c Carrier (Output part of constant reduction planetary gear set)
21d Ring gear 24 4th planetary gear set (subsequent planetary gear set)
24a Sun gear (input part of the following planetary gear set)
27 Pinion shaft 27a Pinion lubricating oil passage 28 Sleeve 60 Hydraulic servo 61 Clutch piston 61a Piston side notch 61b Cylindrical portion of clutch piston 67 Balance hydraulic chamber 71 Seal plate 80 Hub portion 81b Hub side fitting portion 82 Friction plate (multi-plate clutch) )
83 Disc plate (multi-plate clutch)
85 Second lubrication hole 87 First lubrication hole 88 Clutch lubrication oil passage 89 Drum

Claims (6)

An input shaft;
A constant deceleration planetary gear set that constantly decelerates and outputs the rotation of the input shaft;
A subsequent planetary gear set that receives the output rotation of the constant speed planetary gear set;
A multi-plate clutch provided between the output part of the constantly reducing planetary gear set and the input part of the subsequent planetary gear set;
A clutch piston for pressing the multi-plate clutch;
A hydraulic servo for controlling the clutch piston ;
A boss that is fixed to the transmission case and extends in the input shaft direction and through which the input shaft is inserted ,
A hub portion provided on the inner peripheral side of the multi-plate clutch is connected to the output portion of the constant speed planetary gear set, and a drum provided on the outer peripheral side of the multi-plate clutch is the input of the subsequent planetary gear set. An automatic transmission connected to a section,
The clutch piston and the hydraulic servo are provided on the hub portion or a member connected to the hub portion ,
The constantly reducing planetary gear set is a ring gear input and a carrier output, and a sun gear is fixed to the boss part,
The hub part is in communication with the carrier and in contact with a sleeve provided rotatably on the boss part,
The automatic transmission , wherein the clutch piston is in sliding contact with an inner peripheral surface of the hub portion and an outer peripheral surface of the sleeve . The hub portion is notched at a plurality of locations in the circumferential direction of the portion facing the piston, and a hub side fitting portion that is a notch remaining portion is formed,
The clutch piston is formed with a piston-side notch in which a portion corresponding to the hub-side fitting portion is notched,
2. The automatic transmission according to claim 1, wherein the hub portion and the clutch piston are axially overlapped so that the hub-side fitting portion and the piston-side notch portion are fitted. Machine. The clutch piston is formed with a cylindrical portion having an inner peripheral surface that is coaxial with the sleeve outer peripheral surface and larger in diameter than the sleeve outer peripheral surface,
The clutch piston is coaxially adjacent, and the inner peripheral portion is locked to the outer peripheral surface of the sleeve so as not to move away from the clutch piston, and the outer peripheral portion is the inner periphery of the cylindrical portion of the clutch piston. It has a seal plate that slides on the surface,
The control hydraulic chamber of the hydraulic servo is formed by the mutually opposing surfaces of the seal plate and the clutch piston, the inner peripheral surface of the cylindrical portion of the clutch piston, and the outer peripheral surface of the sleeve. The automatic transmission according to claim 1 or 2, characterized in that: The hydraulic servo of the balance hydraulic chamber, claim 3, wherein the opposing surfaces of the clutch piston and the carrier, and the inner peripheral surface of the hub portion, that are formed in the outer peripheral surface of the sleeve Automatic transmission as described. The carrier supports a pinion shaft that supports the pinion of the constant speed planetary gear set,
The pinion shaft is formed with a pinion lubricating oil passage for lubricating the pinion,
5. The automatic transmission according to claim 4, wherein a portion of the carrier forming a balance hydraulic chamber communicates with the pinion lubricating oil passage. A clutch lubricating oil passage is formed between the outer peripheral surface of the hub portion and the inner peripheral surface with which the clutch piston is slidably contacted, and extends in the direction in which the multi-plate clutch is arranged.
A first lubricating hole that communicates a portion of the inner peripheral surface of the hub portion that does not form the balance hydraulic chamber and the clutch lubricating oil passage, and a second lubricating passage that communicates the clutch lubricating oil passage and the outer peripheral surface of the hub portion. The automatic transmission according to any one of claims 1 to 5, wherein a lubrication hole is provided.

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