JP2000055152A - Automatic transmission for vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide an automatic transmission for vehicle of a gearing configuration which can achieve a six-speed operation, whereby the easiness in the on-board mounting works is enhanced by shortening the axial length while the gear noise is reduced. SOLUTION: The transmitting mechanism of an automatic transmission is composed of an input shaft 11, a planetary gear set G having four transmitting elements S2, S3, C2(C3), R3(R2), a decelerating planetary gear G1, two brakes B-1, B-2 and three clutches C-1, C-2, C-3, and a counter-drive gear 19. The decelerating planetary gear G1 is configured on the double pinion type planetary gearing system, wherein the tail of its carrier C1 is coupled with the input shaft 11 while the forefront coupled with the clutch C-2, and a ring gear R1 is coupled with an intermediate shaft 13 installed behind the input shaft 11, and a sun gear S1 is fixed to the forefront of a transmission case 10 and located at the forefront of the transmitting mechanism. A frictional member 55 of the clutch C-2 is located on the periphery of the decelerating planetary gear G1 while the counter-drive gear 19 located behind it G1, and the clutches C-1 and C-3 are coupled with the ring gear R1 of the decelerating planetary gear G1 through the intermediate shaft 13.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an automatic transmission mounted on a vehicle, and more particularly to an arrangement of transmission components in a gear train.

[0002]

2. Description of the Related Art As one form of an automatic transmission for a vehicle, there is a horizontal automatic transmission for a front engine / front drive (FF) vehicle or a rear engine / rear drive (RR) vehicle. In such a type of automatic transmission, the engine and the automatic transmission are arranged in series between the left and right wheels of the vehicle, so that the axial length of the automatic transmission is significantly limited. Therefore, the gear train of such an automatic transmission must have a configuration that minimizes the number of planetary gear sets, which have a large number of shifting elements that mainly extend the shaft length, and the number of clutches and brakes that operate the shifting elements, as much as possible. No.

On the other hand, there is a demand for an automatic transmission having multiple stages in order to not only ensure drivability but also to improve fuel efficiency which is indispensable for energy saving. It is necessary to further reduce the number of transmission elements and the number of clutches and brakes. Therefore,
Japanese Patent Application Laid-Open No. 4-219553 proposes a gear train that uses a planetary gear set having a minimum number of shift elements and achieves six forward speeds and one reverse speed by using three clutches and two brakes for operating the planetary gear set. I have. The gear train according to this proposal employs three clutches to convert the engine output rotation and its reduced rotation to a planetary gear set consisting of four transmission elements of a transmission as appropriate.
The two speeds are input as different speeds, and the two shift elements are locked and controlled by the two brakes to achieve the multi-stage six speeds.

[0004]

As described above, the gear train configuration according to this proposal is very reasonable in the number of shift elements per shift speed and the number of clutches and brakes required. This is not without its practical problems. In general, in a gear that transmits torque by meshing teeth, it is difficult to avoid the occurrence of gear noise during torque transmission. Therefore, a gear that constantly transmits torque, such as a counter drive gear, is provided at the rear end of a transmission (through the specification, The position on the shaft is defined with the power input side as the front). Avoid placing on the side of the vehicle, and place it as close to the center of the vehicle as possible in the width direction, that is, at the front end of the transmission, so that the engine and the like are shielded. It is preferred that objects make noise less likely to leak out of the vehicle (throughout this specification, blocking such noise is referred to as noise reduction). In particular, in the gear train according to the above proposal, the planetary gear for deceleration is in a torque transmitting state in a wide speed range from the first speed to the fifth speed, so that special measures against this gear noise are required. In this regard, in the above-described conventional technology, the clutch and the brake are collectively arranged on the front end side in the axial direction of the planetary gear set, and the counter drive gear is arranged on the rear end portion. There's a problem.

Therefore, it is recalled that the reduction planetary gear and the counter drive gear are arranged on the front side of the speed change mechanism. By the way, in the above-described gear train, although the minimum number of components is used, the number of members increases and the size of the transmission increases as the transmission mechanism shifts to the sixth speed. Therefore, when reducing the gear noise, the countermeasures must not involve an increase in the size of the transmission that impairs the mountability on the vehicle, and in particular, an increase in the total length that causes interference with the vehicle-side member. . Therefore, from this perspective,
The measures against the noise, that is, the arrangement of the reduction planetary gear and the counter drive gear on the front side of the transmission mechanism should not be incompatible with the radial arrangement of the transmission components for shortening the shaft length of the transmission.

SUMMARY OF THE INVENTION The present invention has been made in view of the above circumstances, and provides an automatic transmission for a vehicle in which the gear noise is reduced, the shaft length is not increased due to the increase in the number of gears, and the mountability of the vehicle is improved. With the goal.

[0007]

SUMMARY OF THE INVENTION In order to achieve the above object, the present invention provides, as a transmission mechanism, an input shaft, a planetary gear set having four transmission elements, a reduction planetary gear, two brakes, An automatic transmission for a vehicle in which a clutch and a counter drive gear are arranged, wherein a first transmission element of a planetary gear set is connected to an input shaft via a reduction planetary gear by a first clutch, and a second transmission element is provided. Are connected to the input shaft via a deceleration planetary gear by a third clutch, can be locked to the transmission case by the first brake, and the third transmission element is connected to the input shaft by the second clutch. With the second brake being capable of being locked to the transmission case by the second brake, and the fourth transmission element being connected to the counter drive gear. The reduction planetary gear has a double pinion planetary gear configuration, in which the rear end of the carrier is connected to the input shaft, the front end is connected to the second clutch, and the ring gear is connected to an intermediate shaft provided behind the input shaft. The sun gear is fixed to the front end of the transmission case and is disposed on the front end side of the transmission mechanism. The friction member of the second clutch is disposed on the outer peripheral side of the reduction planetary gear. The counter drive gear is located behind the reduction planetary gear. The first and third clutches are connected to a ring gear of a reduction planetary gear via an intermediate shaft.

In order to further shorten the shaft length, it is effective that the second brake has a band brake structure and is arranged on the outer periphery of the second clutch.

In order to make the rear end of the transmission more compact, the planetary gear set is disposed on one side of the counter drive gear, and the friction member of the first clutch and the friction member of the third clutch are It is effective to adopt a configuration in which the hydraulic servos of the first clutch and the third clutch are arranged on the outer peripheral side of the planetary gear set and are arranged side by side on one side of the planetary gear set.

[0010] Further, in order to achieve overall compactness, a first brake having a band brake structure is arranged on the outer peripheral side of the friction member of the first clutch and the friction member of the third clutch. Is effective.

[0011]

According to the structure of the first aspect,
Since the reduction planetary gear and the counter drive gear are arranged close to the front end side of the transmission mechanism, that is, close to the center of the vehicle, gear noise can be reduced. The reduction planetary gear has a double pinion gear structure, and the rotation from the input shaft is transmitted to the second clutch through the carrier, so that the friction member of the second clutch can be arranged on the outer periphery of the reduction planetary gear. , The overall length can be reduced. At this time, since the second clutch is a relatively small-capacity clutch that does not transmit the deceleration torque, the radial dimension can be relatively small, and interference with the differential ring gear or the like can be prevented at the front end. it can. Also, with the front end arrangement of the second clutch and the deceleration planetary gear, members that need to transmit the input rotation are no longer behind these clutches and gears, so the input shaft is terminated at these arrangement positions, An intermediate shaft is provided to transmit the decelerated rotation from the deceleration planetary gear to the rear, thereby avoiding unnecessary multi-shaft structures, which is advantageous for the overlapping arrangement of each member while suppressing an increase in radial dimension. can do.

According to the second aspect of the present invention, the second brake is configured as a band brake, so that the overall length can be reduced by further overlapping the arrangement while minimizing the increase in the radial dimension. it can.

[0013] Further, according to the structure of the third aspect, the rear end portion can be made compact in the radial direction, and as a whole,
A compact layout can be realized.

According to the fourth aspect of the present invention, a band brake that is long in the axial direction and compact in the radial direction is disposed on the outer periphery of the friction member of the first and third clutches, so that space can be efficiently used. The structure can be made compact in the radial direction, and the weight can be reduced. Furthermore, when the valve body is mounted in front of the automatic transmission,
Interference with auxiliary equipment such as a radiator fan can be avoided, and a crushable zone can be secured. In addition, when mounted below, a larger minimum ground clearance can be secured.

[0015]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 shows a gear train according to a first embodiment of a vehicular automatic transmission embodying the present invention, which is developed by a skeleton with a shaft-to-shaft developed in a common plane. FIG. 2 shows an actual axial positional relationship of the automatic transmission as viewed from an end face. This automatic transmission has a main axis X, a counter axis Y,
It has a three-axis configuration in which each element is disposed on each axis of the differential axis Z. Then, as a transmission mechanism, the input shaft 11, a planetary gear set G having four transmission elements, a reduction planetary gear G1, and two brakes B-1 and B-2.
, Three clutches C- 1, C- 2 and C- 3 and a counter drive gear 19 are arranged on the main shaft X.

In this automatic transmission, the first transmission element S3 of the planetary gear set G is connected to the input shaft 11 via the reduction planetary gear G1 by the first clutch C-1, and the second transmission element S2 is connected to the third transmission element S3. Is connected to the input shaft 11 via the deceleration planetary gear G1 by the clutch C-3, and can be locked to the transmission case 10 by the first brake B-1, and the third transmission element C2 (C3) The second clutch B-2 is connected to the input shaft 11 and the second brake B-2 is capable of being locked to the transmission case 10. The fourth transmission element R2 (R3) is an output element on the main shaft X. As a counter drive gear 19. In the gear train shown in the figure, a one-way clutch F-1 is arranged in parallel with the brake B-2. This is because a one-way clutch F-2 and a brake B- In order to avoid complicated hydraulic control for re-grip 1 and to simplify the release control of the brake B-2, a one-way clutch F-1 that releases the engagement force naturally with the engagement of the brake B-1 is used. And is equivalent to the brake B-2.

Hereinafter, the gear train of this embodiment will be described in more detail. A torque converter 4 with a lock-up clutch for transmitting the rotation of an engine (not shown) to the input shaft 11 is arranged on the main shaft X. A counter gear 2 is arranged on the counter shaft Y. Canter gear 2
Is provided with a large diameter cantilever driven gear 21 fixed to the counter shaft 20 and meshing with the counter drive gear 19, and a small diameter differential drive pinion gear 22 similarly fixed to the counter shaft 20 and meshing with the differential ring gear 31. These functions reduce the output from the main shaft X side, and reversely transmit the output to the differential device 3. A differential device 3 is provided on the differential shaft Z. Differential device 3
Is configured such that a differential case 32 is provided fixed to a differential ring gear 31 and the differential rotation of a differential gear disposed therein is output to the left and right shafts 30 to be the final wheel driving force. I have.

The planetary gear set G is meshed with a pair of large and small sun gears S2 and S3, one of which meshes with the large diameter sun gear S2 and the ring gear R2.
(R3) and a carrier C supporting a pair of pinion gears P2 and P3, the other of which meshes with the small-diameter sun gear S3.
It is composed of a Ravigneaux type gear set composed of C2 and C3. In this embodiment, the small-diameter sun gear S3 is the first transmission element, the large-diameter sun gear S2 is the second transmission element,
The carriers C2 and C3 serve as a third speed change element, and the ring gear R2 (R3) serves as a fourth speed change element.

The reduction planetary gear G1 is fixed to the transmission case 10 using the sun gear S1 as a reaction element, connected to the input shaft 11 using the carrier C1 as an input element, and connected to the first clutch C-1 using the ring gear R1 as an output element. And a planetary gear set G via a third clutch C-3. The first transmission element of the planetary gear set G, that is, the small-diameter sun gear S3, is connected to the first clutch C-1, and the second transmission element, that is, the large-diameter sun gear S2, is connected to the third clutch C-3. And the first brake B-
1 allows the automatic transmission case 10 to be locked. Also, the carrier C2 (C3), which is the third speed change element
Is connected to the input shaft 11 via the second clutch C-2, and is connected to the transmission case 1 by the second brake B-2.
0 and the one-way clutch F
-1 allows the transmission case 10 to be rotationally locked in one direction. The fourth speed change element, that is, the ring gear R2 (R3) is connected to the counter drive gear 19.

The automatic transmission having such a configuration shifts gears based on the vehicle load and the vehicle speed in a speed range corresponding to the range selected by the driver under the control of an electronic control unit and a hydraulic control unit (not shown). Do. FIG. 3 graphically illustrates the shift speeds achieved by engaging and disengaging the clutches and brakes (showing engagement with a circle and disengagement without a mark). Also,
FIG. 4 is a speed diagram showing the relationship between the shift speed achieved by the engagement of each clutch and brake (the engagement is indicated by a mark) and the rotational speed ratio of each shift element at that time.

As can be seen by referring to both figures, the first speed (1ST) corresponds to the engagement of the clutch C-1 and the brake B-2 (in this embodiment, as shown in the operation table, One-way clutch F instead of engagement of brake B-2
The automatic engagement of -1 is used. The reason for using this engagement and the reason why this engagement corresponds to the engagement of the brake B-2 will be described later in detail. ).
In this case, the rotation reduced from the input shaft 11 via the reduction planetary gear G1 is input to the small-diameter sun gear S3 via the clutch C-1, and a reaction force is applied to the carrier C3 locked by the engagement of the one-way clutch F-1. The reduction rotation of the ring gear R3 at the maximum reduction ratio is the counter drive gear 1
9 is output.

Next, in the second speed (2ND), the clutch C-
This is achieved by the engagement of 1 with the brake B-1. In this case, the rotation reduced from the input shaft 11 via the reduction planetary gear G1 is input to the small-diameter sun gear S3 via the clutch C-1, and the reaction force is applied to the large-diameter sun gear S2 locked by the engagement of the brake B-1. Take ring gear R2 (R3)
Is output to the counter drive gear 19.
As shown in FIG. 4, the reduction ratio at this time is the first speed (1S
T).

In the third speed (3RD), the clutch C-
1 and the clutch C-3 at the same time. In this case, the rotation reduced from the input shaft 11 via the reduction planetary gear G1 is simultaneously input to the large-diameter sun gear S2 and the small-diameter sun gear S3 via the clutch C-1 and the clutch C-3, so that the planetary gear set G is directly connected. The rotation of the ring gear R2 (R3), which is the same as the input rotation of both sun gears, is output to the counter drive gear 19 as a rotation that is reduced with respect to the rotation of the input shaft 11.

Further, in the fourth speed (4TH), the clutch C-
1 and the clutch C-2 at the same time. In this case, on the other hand, the rotation reduced from the input shaft 11 via the reduction planetary gear G1 is transmitted to the sun gear S via the clutch C-1.
3, the non-decelerated rotation input from the input shaft 11 via the clutch C-2 is input to the carrier C3, and a rotation intermediate between the two input rotations is The rotation is output to the counter drive gear 19 as the slightly reduced rotation of the ring gear R3.

Next, in the fifth speed (5TH), the clutch C-
2 and the clutch C-3 at the same time. In this case, on the other hand, the rotation reduced from the input shaft 11 via the reduction planetary gear G1 is transmitted to the sun gear S via the clutch C-3.
2, the non-decelerated rotation input from the input shaft 11 via the clutch C-2 is input to the carrier C2, and the rotation slightly higher than the rotation of the input shaft 11 of the ring gear R2 is used as the counter drive. Output to the gear 19.

In the sixth speed (6TH), the clutch C
-2 and brake B-1. In this case, the non-decelerated rotation is input only from the input shaft 11 to the carrier C2 via the clutch C-2, and the brake B-1
The further increased rotation of the ring gear R2 that takes a reaction force on the sun gear S2 locked by the engagement of is output to the counter drive gear 19.

The reverse (REV) is controlled by the clutch C-3.
And the brake B-2. in this case,
The rotation reduced from the input shaft 11 via the reduction planetary gear G1 is input to the sun gear S2 via the clutch C-3, and the carrier C locked by the engagement of the brake B-2.
2 is output to the counter drive gear 19.

Here, the one-way clutch F mentioned earlier
The relationship between -1 and the brake B-2 will be described. As can be seen from the engagement / disengagement relationship between the two brakes B-1 and B-2 at the first speed and the second speed, these two brakes release one of them during the up-down shift between the two shift speeds. At the same time, the other engagement is performed, that is, a so-called gripped friction element. Such gripping of frictional elements requires precise simultaneous control of the engagement pressure and release pressure of the hydraulic servos that operate them, and such control requires the addition of control valves and complicated hydraulic circuits. Will be invited. Therefore, in the present embodiment, the engagement direction of the one-way clutch F-1 is changed to the first speed by utilizing the fact that the reaction torque applied to the carrier C2 (C3) is reversed between the first speed and the second speed. By setting according to the reaction torque supporting direction,
By causing the one-way clutch F-1 to exhibit substantially the same function as the engagement of the brake B-2, the brake B-
2 (in the vehicle coast state of the wheel drive, the direction of the reaction torque applied to the carrier C2 (C3) is reversed with respect to the state of the engine drive. Therefore, in order to obtain the engine braking effect, ○ in parentheses in Figure 3
Requires the engagement of brake B-2, as indicated by the mark),
That is, the carrier C2 (C3) is locked.
Therefore, in order to achieve the shift speed, a configuration in which the first speed is achieved by engaging the brake B-2 without providing a one-way clutch may be adopted.

Each of the speed stages achieved in this manner can be qualitatively understood on the speed diagram of FIG. 4 by referring to the vertical interval between the circles indicating the speed ratio of the ring gears R2 and R3. Thus, good speed steps with relatively equal intervals for each shift speed are obtained. When this relationship is specifically set to numerical values and quantitatively expressed, the gear ratio shown in FIG. 3 is obtained. The gear ratio in this case is such that the gear ratio λ1 = 0.333 between the sun gear S1 of the reduction planetary gear G1 and the ring gear R1, the sun gear S2 and the ring gear R2 on the large diameter sun gear side of the planetary gear set G.
(R3) tooth number ratio λ2 = 0.436, tooth number ratio λ3 = 0.35 between sun gear S3 and ring gear R3 on the small diameter sun gear side
When it is set to 9, the input / output gear ratio and the steps are as shown in FIG. 3, and the gear ratio width is 6.000.

Next, FIG. 5 is a schematic cross-sectional view showing the structure of the automatic transmission more specifically. The components described above with reference to the skeleton will be denoted by the same reference numerals and will not be described, but details that cannot be referenced from the skeleton will be described here. According to a feature of the present invention, the reduction planetary gear G1 has a double pinion planetary gear configuration including a carrier C1 that supports a pair of pinion gears P1 and P1 ′ that mesh with each other, one of which meshes with the sun gear S1, and the other meshes with the ring gear R1, The rear end of the carrier C1 is connected to the input shaft 11, and the front end is connected to the second clutch C-2.
Is connected to an intermediate shaft 13 provided at the rear of the input shaft 11, and the sun gear S1 having the front side opened in this manner is connected to the boss 10a extending rearward from the front end of the transmission case 10.
And is disposed on the front end side of the transmission mechanism. Further, the friction member 55 of the second clutch C-2 is arranged on the outer peripheral side of the reduction planetary gear G1. Further, the counter drive gear 19 is arranged behind the reduction planetary gear G1. Further, the first clutch C-1 and the third clutch C-3 are connected via an intermediate shaft 13 to the ring gear R1 of the reduction planetary gear G1.

The planetary gear set G is disposed on one side of the counter drive gear 19 and has a first clutch C-
Hydraulic servos 6 and 7 of the first and third clutches C-3 are arranged side by side on one side of the planetary gear set G, and friction members 6 of the first clutch C-1 and the third clutch C-3 are arranged.
The reference numerals 3 and 73 are arranged from the outer peripheral side of the hydraulic servo 6 to the outer peripheral side of the planetary gear set G. Furthermore, the first
Friction members 6 of the first clutch C-1 and the third clutch C-3
A first brake B-1 having a band brake configuration is arranged on the outer peripheral side of the third and third brakes 73. By arranging the two clutches C-1 and C-3 in the axial direction in this manner,
Both clutches have approximately the same capacity. Also, on the outer periphery of these clutches which are arranged side by side and have the same outer diameter,
By arranging a band brake that is long in the axial direction and compact in the radial direction, there is no interference with the differential device 3 and the vehicle-side member B in the automatic transmission, so that the radial dimension restriction is relatively gentle, so that the intermediate portion of the transmission is relatively gentle. The space in the axial direction is reduced without wasting space, thereby reducing the axial dimension.

In this embodiment, the hydraulic servo 6 for the first clutch C-1 and the hydraulic servo 7 for the third clutch C-3
The torque transmission members 60 connected to the intermediate shaft 13 connected to the carrier C1 of the reduction planetary gear G1 are configured as respective hydraulic servo cylinders. Specifically, the torque transmission member 60 has a front end connected to the intermediate shaft 13, a first cylindrical portion 60 a supported by a boss 10 b extending forward from the rear end of the transmission case 10, The first cylindrical portion 60a extends in the outer diameter direction from a substantially central portion in the axial direction,
A second cylindrical portion 60b formed into a cylindrical shape by bending forward and backward at a predetermined diameter, and a first clutch C
-1 and the drum portion 6 extending toward the third clutch C-3
The hydraulic servo cylinders of the first clutch C-1 and the third clutch C-3 are configured to be back-to-back with the two cylindrical portions 60a and the portion extending in the outer diameter direction, respectively.

The two hydraulic servos 6,
7 pistons 61 and 71 are inserted. Accordingly, the hydraulic servo 6 employs a normal servo format in which the drum portion 62 side integral with the torque transmitting member 60 is a cylinder, and the piston 61 is inserted into the cylinder. This is a servo type reverse to that of a normal type, in which the drum unit 62 side integral with the piston is an axially immovable piston, and the drum unit 72 extending from the piston 71 is an axially movable cylinder. From the relationship of the servo type, the friction member of the clutch C-1 (comprising a disk and a separator plate; the same applies to other clutches) 63
The drum portion 72 of the clutch C-3 that covers the outside of the drum portion 62 that supports the outer periphery of the drum portion 62 of the clutch C-1
The drum member 72 supports the outer periphery of the friction member 73 by spline engagement. With this arrangement, the decelerated rotation entering the torque transmitting member 60 is always performed by the drum portion 62 of the clutch C-1 and the clutch C
-3 is transmitted to the drum section 72. Therefore, when the piston 61 is pushed out of the cylinder by the operation of the hydraulic servo 6, the axially immovable drum portion 62 and the piston 61
And the friction member 63 is sandwiched between the
4 to the sun gear S3. On the other hand, when the piston 71 is pushed back with respect to the drum portion 62 by the operation of the hydraulic servo 7, the friction member 73 is sandwiched between the drum portion 72 and the drum portion 62 that is not moved in the axial direction, so that the deceleration rotation is performed. The power is transmitted to the sun gear S2 via the hub 74. In FIG. 5, reference numeral 65 denotes a cancel plate that cancels the centrifugal oil pressure of the hydraulic servo 6, and reference numeral 75 similarly denotes a cancel plate that cancels the centrifugal oil pressure of the hydraulic servo 7.

With respect to the hydraulic servo structure of the second clutch C-2, in this embodiment, the hydraulic servo 5 for operating the clutch C-2 has a cylinder and a piston built in the front end of the transmission case 10. It is a stationary cylinder type hydraulic servo. Specifically, the cylinder 50 is formed in the transmission case 10 as an annular groove, and inside thereof,
Similarly, a ring-shaped piston 51 is fitted slidably in the axial direction. And this piston 51
Pressure plate 5 through thrust bearing 52
3, the clutch friction member 55 is held between the drum 56 connected to the carriers C2 and C3 of the planetary gear set G, and the input rotation from the hub connected to the carrier C1 is transmitted via the drum 56. Carrier C
2, C3.

With respect to the support of the counter drive gear 19, in this embodiment, the counter drive gear 19
The outer periphery of the axially extending portion of the hub is supported by the inner periphery of a support 10c of the case 10 also serving as a hydraulic servo cylinder of the second brake B-2 via a bearing 12.

In this embodiment, the brake B-2 is configured such that the friction member 84 having a multi-plate structure is engaged by the hydraulic servo 8 built in the support 10c, and is disposed adjacent to the friction member 84. It is connected to the carrier C2 via the inner race 9 of the one-way clutch F-1. The brake B-2 and the one-way clutch F-1 are arranged on the outer peripheral side of the planetary gear set G.

Thus, in this embodiment, since the reduction planetary gear G1 and the counter drive gear 19 are arranged close to the front end side of the transmission mechanism, that is, close to the center of the vehicle, gear noise can be reduced. The reduction planetary gear G1 is configured as a double pinion gear, and the rotation from the input shaft 11 is transmitted to the second clutch C-2 through the carrier C1.
By arranging the friction member 55 of the second clutch C-2 on the outer periphery of, the overall length is shortened by the overlapping arrangement of the members. In this case, the second clutch C-2 is a relatively small-capacity clutch that does not transmit the deceleration torque, so that the radial dimension can be relatively small, and interference with the differential ring gear 31 and the like at the front end is prevented. Have been. Further, with the front end arrangement of the second clutch C-2 and the reduction planetary gear G1, there is no member that needs to transmit the input rotation behind the clutch and the gear. And an additional intermediate shaft 13 is provided to transmit the decelerated rotation from the deceleration planetary gear G1 rearward, thereby avoiding an unnecessary multi-shaft structure. The increase in the radial dimension is suppressed irrespective of the overlapping arrangement of the friction member 73, the friction member 84 of the first brake, the one-way clutch F-1 and the second brake.

Further, by arranging the hydraulic servos 6 and 7 at the rear end, it is possible to reduce the size of the rear end in the radial direction, and to realize a compact layout as a whole. Also, the first and third clutches C-1,
Along the circumference of the friction members 63 and 73 of C-3,
By arranging the compact band brake B-1 in the radial direction, the space can be used without waste and the configuration is compact in the radial direction.

In the first embodiment, the planetary gear set G is a Ravigneaux type. However, the gear set G that can take a relatively good gear ratio and steps is not limited to this. Therefore, an embodiment in which the planetary gear set G is changed to another type will be described below.

FIG. 6 shows a second embodiment in which only the planetary gear set G is partially changed from the first embodiment. In this embodiment, the planetary gear set G is a combination of a simple planetary gear G2 and a double pinion type planetary gear G3 in which one of a pair of pinion gears P3 and P3 'meshing with each other meshes with a sun gear S3 and the other meshes with a ring gear R3. It has a configuration. In the case of this planetary gear set G, the first clutch C-
1 is connected to the two sun gears S2 and S3, the third clutch C-3 is connected to the ring gear R2 of the simple planetary gear G2, and the second clutch C-2 is connected to the carrier C2.
And the carrier C3, and the ring gear R3 is connected to the counter drive gear 19. The brake B-1 is a ring gear R2 of the simple planetary gear G2.
, And the brake B-2 and the one-way clutch F-1 lock both the carrier C2 and the carrier C3. Therefore, in this embodiment, both sun gears S2 and S3 are the first transmission element, ring gear R2 is the second transmission element, carriers C2 and C3 are the third transmission element, and ring gear R3 is the fourth transmission element. I have.
In such a case, for example, a gear ratio and a step as shown in Table 1 below are obtained.

[Table 1] Incidentally, in this case, the gear ratio λ1 of the reduction planetary gear G1 = 0.333, and the gear ratio λ2 of the large-diameter ring gear side =
0.636, tooth ratio λ3 = 0.33 on the small diameter ring gear side
3, and the gear ratio width is 6.857.

FIG. 7 shows a third embodiment in which the positional relationship between the simple planetary gears and the double planetary gears of the planetary gear set G is reversed with respect to the second embodiment. In this embodiment, the first clutch C-1 is connected to the small-diameter sun gear S2 on the double planetary gear G2 side, and the third clutch C-3 is connected to the carrier C2 of the double planetary gear G2 and the large-diameter sun gear S3 of the simple planetary gear G3. The second clutch C-2 is connected to the carrier C3 of the simple planetary gear G3 and the ring gear R2 of the double planetary gear G2, and the ring gear R3 of the simple planetary gear G3 is connected to the counter drive gear 1.
9. The brake B-1 locks the carrier C2 and the large-diameter sun gear S3, and the brake B-2 and the one-way clutch F-1 lock the ring gear R2 and the carrier C3. Therefore, in this embodiment, the small-diameter sun gear S2 is the first transmission element, the carrier C2 and the large-diameter sun gear S3 are the second transmission element, the ring gear R2 and the carrier C3 are the third transmission element, and the ring gear R3 is the fourth transmission element. It is a speed change element.
In such a case, for example, a gear ratio and a step as shown in Table 2 below are obtained.

[Table 2] Incidentally, in this case, the gear ratio of the reduction planetary gear G1 is λ1 = 0.333, the gear ratio of the double planetary gear G2 is λ2 = 0.474, the gear ratio of the simple planetary gear G3 is λ3 = 0.444, and the gear ratio width is 6.022.

Next, FIG. 8 shows the planetary gear set G
A fourth embodiment composed of two double planetary gears G2, G3 is shown. In this embodiment, the first clutch C-1 is connected to the sun gear S3 and the carrier C2, the third clutch C-3 is connected to the sun gear S2, and the second clutch C
-2 is connected to the carrier C3 and the ring gear R2, and the ring gear R3 is connected to the counter drive gear 19. The brake B-1 locks the sun gear S2, and the brake B-2 and the one-way clutch F-
Reference numeral 1 is used to lock the ring gear R2 and the carrier C3. Therefore, in this embodiment, the small diameter sun gear S3 and the carrier C2 are the first speed change element, the large diameter sun gear S2 is the second speed change element, and the ring gear R2 and the carrier C3 are the third speed change element.
And the ring gear R3 is a fourth speed change element. In such a case, for example, a gear ratio and a step as shown in Table 3 below are obtained.

[Table 3] Incidentally, in this case, the gear ratio λ1 of the reduction planetary gear G1 = 0.333, and the gear ratio λ2 of the planetary gear G2 =
0.556, gear ratio λ3 = 0.3 of planetary gear G3
33, and the gear ratio width is 6.375.

Next, FIG. 9 shows a fifth embodiment in which only the connection relationship is changed from the fourth embodiment. In this form,
The first clutch C-1 is connected to both sun gears S2 and S3, the third clutch C-3 is connected to the carrier C2,
The second clutch C-2 includes a carrier C3 and a ring gear R2.
And the ring gear R3 is connected to the counter drive gear 1
9. The brake B-1 locks the carrier C2, and the brake B-2 and the one-way clutch F-1 lock the ring gear R2 and the carrier C3. In this embodiment, both the large-diameter and small-diameter sun gears S2 and S3 are provided by the first transmission element, the carrier C2
Are the second speed change element, the ring gear R2 and the carrier C3 are the third speed change element, and the ring gear R3 is the fourth speed change element. In such a case, for example, a gear ratio and a step as shown in Table 4 below are obtained.

[Table 4] Incidentally, in this case, the gear ratio λ1 of the reduction planetary gear G1 = 0.333, and the gear ratio λ2 of the planetary gear G2 =
0.444, gear ratio λ3 of planetary gear G3 = 0.3
61, and the gear ratio width is 6.029.

Each of the above embodiments (in the following description,
These are collectively referred to as a first group of embodiments), in which the planetary gear set G is changed on the basis of the first embodiment, and then, the arrangement of the second brake B-2 is changed. A description will be given by giving an example.

In the sixth embodiment shown in a schematic cross section of the skeleton in FIGS. 10 and 11, the second brake B-2
Is a band brake configuration, and the second clutch C-2
It is arranged on the outer periphery of. In this embodiment, with the band brake of the second brake, the support structure of the canter drive gear 19, the first and third clutches C-1, C-3, and the one-way Since the arrangement position of the clutch F-1 has been changed, only the relevant portion will be described.

First, the second brake B-2 is connected to the second clutch C at the same axial position as the differential ring gear 31.
-2 as a band brake. The overlapping arrangement of the second clutch C-2 and the band brake B-2 in the radial direction allows the clutch and the brake to be arranged at the same axial position while avoiding an increase in the radial dimension that causes interference with the differential ring gear 31. It is an effective method for

The first and second clutches C-1, C-3 and the one-way clutch F-1 are utilized by utilizing the empty space occupied by the multi-plate friction member in the first embodiment. It is disposed substantially on the outer peripheral side of the planetary gear set G so as to be entirely forward. Since such an arrangement is useful for reducing the diameter of the outer peripheral portion of the hydraulic servo installation section, in order to avoid interference with the vehicle side member, the first arrangement is used.
This is an advantageous arrangement over the group embodiment.

Also in this embodiment, the hydraulic servo 6 of the first clutch C-1 and the hydraulic servo 7 of the third clutch C-3 are torque transmitting members connected to the intermediate shaft 13 connected to the carrier C1 of the reduction planetary gear G1. Although 60 is configured as each hydraulic servo cylinder, the relationship between the torque transmitting member 60 and the piston 71 is slightly different from that of the first embodiment. Specifically, the second cylindrical portion is not formed by bending the torque transmitting member 60 forward and backward, and the torque transmitting member 60 is connected to the first clutch C-1 and the third clutch 3 from a predetermined diameter. The piston 71 is connected to the drum portion 62 extending toward the clutch C-3.

In this embodiment, the support structure of the case 10
A boss is formed by extending the inner peripheral side of Oc to the rear in the axial direction,
A configuration is employed in which the inner circumference of the counter drive gear 19 is supported on the outer circumference via a bearing 12. The advantage of this configuration is that the load due to the moment force applied to the bearing 12 can be reduced by supporting the inner peripheral side of the meshing portion of the counter drive gear 19 with the driven gear 21 without offset in the axial direction. The point is that a bearing with a small capacity can be obtained.

In the automatic transmission having this arrangement, the second brake B-2 is configured as a band brake,
While minimizing the increase in the radial dimension, the overall length is shortened by the overlapping arrangement with the second clutch C-2.

For this gear train configuration,
It is possible to change the configuration of the planetary gear and the connection relationship between the transmission element and the clutch and the brake as described in the first embodiment.

FIG. 12 shows a seventh embodiment in which only the planetary gear set G is partially changed from the sixth embodiment. In this embodiment, the configuration of the planetary gear set G and the connection relationship between the respective shift elements and the respective clutches and brakes are the same as those in the second embodiment, and the gear ratios and steps obtained thereby have the same teeth. The same applies to the setting of the numerical ratio.
The description is replaced with the description of the embodiment.

Next, FIG. 13 shows an eighth embodiment using the same configuration of the planetary gear set G as in the third embodiment, and the connection relationship between each speed change element and each clutch and brake. Therefore, also in this case, the obtained gear ratio and step become the same by setting the same gear ratio as in the third embodiment, and these points will be described with reference to the description of the third embodiment. Substitute description.

FIG. 14 shows a ninth embodiment in which the same planetary gear set G as that of the fourth embodiment is used, and the connection relationship between each speed change element and each clutch and brake is used. Also in this case, the obtained gear ratio and step become the same by setting the same gear ratio as in the fourth embodiment, and these points will be described with reference to the description of the fourth embodiment. Substitute.

FIG. 15 shows a tenth embodiment using the same configuration of the planetary gear set G as in the fifth embodiment and the connection relationship between each speed change element and each clutch and brake. In this case, the description is the same as above, and the description is replaced with the description of the fifth embodiment.

As described above, the present invention has been described in detail with reference to the embodiment in which the form and arrangement of the components and the connection relationship are changed.
These are exemplarily limited to those that can provide relatively good gear ratio steps, and the present invention is not limited to these embodiments, but is described in the individual claims of the claims. The present invention can be implemented by changing various specific configurations within the scope of the above items.

[Brief description of the drawings]

FIG. 1 is an expanded skeleton diagram showing a gear train of a first embodiment of a vehicle automatic transmission to which the present invention is applied.

FIG. 2 is an axial end view showing an actual three-axis positional relationship of the gear train.

FIG. 3 is a chart showing the operation of the gear train and the gear ratios and gear ratio steps achieved.

FIG. 4 is a velocity diagram of the gear train.

FIG. 5 is a sectional view schematically showing only a main shaft portion of the gear train.

FIG. 6 is a schematic sectional view of a main shaft portion of a second embodiment in which a planetary gear set of the gear train is changed.

FIG. 7 is a schematic sectional view of a main shaft portion of a third embodiment in which a planetary gear set is similarly changed to another embodiment.

FIG. 8 is a schematic sectional view of a main shaft portion of a fourth embodiment in which a planetary gear set is further changed to another embodiment.

FIG. 9 is a schematic cross-sectional view of a main shaft portion of a fifth embodiment in which a planetary gear set is further changed to another embodiment.

FIG. 10 is a skeleton diagram showing a gear train according to a sixth embodiment in which the type and arrangement of a second brake of the gear train are changed.

FIG. 11 is a schematic sectional view of a main shaft portion of the gear train according to the sixth embodiment.

FIG. 12 is a schematic sectional view of a main shaft portion of a seventh embodiment in which a planetary gear set of the gear train is changed.

FIG. 13 is a schematic sectional view of a main shaft portion of an eighth embodiment in which a planetary gear set is further changed in the same manner.

FIG. 14 is a schematic cross-sectional view of a main shaft portion of a ninth embodiment in which a planetary gear set is further changed.

FIG. 15 is a schematic cross-sectional view of a main shaft portion of a tenth embodiment in which a planetary gear set is further changed.

[Explanation of symbols]

G planetary gear set G1 double pinion planetary gear (reduction planetary gear) S2, S3 sun gear (transmission element) C2, C3 carrier (transmission element) R2, R3 ring gear (transmission element) S1 sun gear C1 carrier R1 ring gear B-1 first brake B- 2 Second brake C-1 First clutch C-2 Second clutch C-3 Third clutch 6,7 Hydraulic servo 10 Transmission case 11 Input shaft 13 Intermediate shaft 19 Counter drive gear

 ──────────────────────────────────────────────────続 き Continuing on the front page (72) Inventor Masahiro Hayabuchi 10 Takane, Fujii-machi, Anjo, Aichi Prefecture Inside Aisin AW Co., Ltd. (72) Satoru Kasuya 10 Takane, Fujii-machi, Anjo, Aichi Prefecture Aisin・ F-term in AW Co., Ltd. (Reference) 3J028 EA01 EA25 EB08 EB13 EB26 EB31 EB37 EB54 EB62 EB66 FA06 FB06 FC13 FC16 FC17 FC20 FC24 FC25 FC63 HA14

Claims (4)

[Claims] An automatic transmission for a vehicle in which an input shaft, a planetary gear set having four transmission elements, a reduction planetary gear, two brakes, three clutches, and a counter drive gear are arranged as a transmission mechanism. Wherein the first transmission element of the planetary gear set is connected to the input shaft via a reduction planetary gear by a first clutch, and the second transmission element is connected to the input shaft via a reduction planetary gear by a third clutch. And a third brake element is connected to the input shaft by a second clutch, and is lockable to the transmission case by a second brake. A fourth speed change element connected to a counter drive gear, wherein the reduction planetary gear is a double pinion planetary gear; As configuration, it is connected to the input shaft of the rear side of the carrier, is connected to the front end side to the second clutch is connected to an intermediate shaft provided at the rear of the input shaft to the ring gear,
The sun gear is fixed to the front end of the transmission case and arranged on the front end side of the transmission mechanism. The friction member of the second clutch is arranged on the outer peripheral side of the reduction planetary gear. The counter drive gear is arranged behind the reduction planetary gear. An automatic transmission for a vehicle, wherein the first clutch and the third clutch are connected to a ring gear of a reduction planetary gear via an intermediate shaft. 2. The automatic transmission according to claim 1, wherein the second brake has a band brake configuration and is arranged on an outer periphery of the second clutch. 3. The planetary gear set is disposed on one side of a counter drive gear. The friction member of the first clutch and the friction member of the third clutch are disposed on the outer peripheral side of the planetary gear set. 4. The automatic transmission for a vehicle according to claim 1, wherein the hydraulic servos of the third clutch and the third clutch are arranged side by side on one side of the planetary gear set. 4. The automatic transmission for a vehicle according to claim 4, wherein a first brake having a band brake structure is disposed on an outer peripheral side of the friction member of the first clutch and the friction member of the third clutch.