JP2009197927A - Multistage transmission planetary gear train - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a multistage transmission planetary gear train of forward ten stages wherein manufacturing cost is reduced, and efficiency of power transmission is increased to enhance fuel economy. <P>SOLUTION: A planetary gear group 18 includes a first planetary gear set 14, a second planetary gear set 16, an input member 42, output members 48, 52, a first reaction member 58, and second reaction members 40, 50. The input shaft 10 is engaged with a first carrier 28 and simultaneously capable of engaging with a second sun gear 30. A first ring gear 22 is capable of engaging with an input member 42, a first sun gear 20 is capable of engaging with the second sun gear 30, and simultaneously is capable of being secured to a stationary part 70, a first reaction member 58 is capable of engaging with a second carrier 38, and simultaneously is capable of being secured to the stationary part 70. The second reaction members 40, 50 are engaged with the second ring gear 32. The output members 48, 52 are engaged with an output shaft 12, and the second planetary gear set 16 can be integrated. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a planetary gear train that can be used for a vehicular automatic transmission and capable of multi-stage shifting.

As planetary gear trains used in vehicle automatic transmissions, gears capable of multi-speed shifting exceeding eight forward speeds have been proposed with the main objective of improving vehicle fuel efficiency, exhaust characteristics, acceleration performance, and the like.
As a conventional planetary gear train capable of such multi-speed shifting, there is a multi-speed planetary gear train using two sets of double pinion type planetary gears, two sets of single pinion type planetary gears and seven friction elements. RU This gear train obtains a forward gear ratio of nine steps by switching so that two of the seven friction elements are always engaged. (See Patent Document 1).

However, since the above conventional planetary gear train requires two pairs of double pinion type planetary gears in order to obtain a preferable gear ratio for an automobile in nine forward speeds, it is manufactured in comparison with a so-called single pinion type planetary gear. In addition to high costs, there are problems that the power transmission efficiency is poor due to the large meshing of the gears. In addition, in the above conventional planetary gear train, five of the seven friction elements are always idle, and the drag torque of these idle friction elements is large. Coupled with the problem of poor power transmission efficiency, there was a problem that the fuel consumption of the car was bad and the heat generation was high.
JP 2007-71374 A

The problem to be solved is that the manufacturing cost is high because the manufacturing cost is high and the power transmission efficiency is low and two sets of double pinion type planetary gears are required, and the five friction elements are always idle. The fuel consumption is bad and the heat generation is high.
An object of the present invention is to increase the ratio of the number of single pinion type planetary gears as compared with the conventional planetary gear train described above, thereby reducing the manufacturing cost while ensuring a preferable gear ratio for a vehicle such as an automobile, and always reducing the number of planetary gears. An object of the present invention is to provide a multi-speed planetary gear train capable of reducing fuel friction and heat generation by reducing the number of rolling friction elements.

  The multi-stage planetary gear train of the present invention includes an input shaft, an output shaft, a first pinion and a first pinion that are arranged coaxially with the input shaft and mesh with the first sun gear, the first ring gear, the first ring gear, and the first sun gear. The first planetary gear set having a first carrier that rotatably supports the shaft, and is disposed between the input shaft and the output shaft, and meshes with the second sun gear, the second ring gear, the second ring gear, and the second sun gear. The second planetary gear set including the second pinion, the second carrier that rotatably supports the second pinion, and the coaxial arrangement with the output shaft, the input member, the output member, and the torque from the input member to the output member The first reaction force member that receives the reaction force torque when transmitting the force, and the reaction force torque that is smaller than the reaction force torque that the first reaction force member receives when transmitting the torque from the input member to the output member A planetary gear group having a second reaction force member to be received, the input shaft is connected to the first carrier and is connectable to the second sun gear, and the first ring gear is connected to the input member. The first sun gear can be connected to the second sun gear and can be fixed to the stationary part, and the first reaction force member can be connected to the second carrier and fixed to the stationary part. The second reaction force member is connected to the second ring gear, the output member is connected to the output shaft, and a clutch capable of integrating the second planetary gear set is provided.

  Since the multi-speed planetary gear train of the present invention is configured as described above, while obtaining a preferable gear ratio for a transmission for an automobile or the like, the ratio of using a single pinion type is increased to reduce the manufacturing cost and improve the power transmission efficiency. It is possible to improve the fuel consumption and heat generation by reducing the number of friction elements that are always free-wheeling to three, which is two less than the conventional example.

  Hereinafter, a multi-speed planetary gear train according to an embodiment of the present invention will be described with reference to the drawings based on examples.

FIG. 1 is a skeleton diagram showing a planetary gear train according to the first embodiment of the present invention.
In the multi-speed planetary gear train of the embodiment shown in FIG. 1, the input shaft 10 and the output shaft 12 driven from the engine 1 via the torque converter 2 are on the same shaft as the output shaft 1a of the engine 1, and the output The shaft 12 drives a drive wheel (not shown).

A first planetary gear set 14, a second planetary gear set 16, a third planetary gear set 18a, and a fourth planetary gear set 18b, which are arranged coaxially with the input shaft 10 and the output shaft 12 from the upstream side toward the downstream side. Are generally called single pinion types, and each has the same configuration.
That is, the first planetary gear set 14 is capable of rotating the first sun gear 20, the first ring gear 22, the plurality of first pinions 24 meshed with the first ring gear 22 and the first sun gear 20, and the first pinion 24. The rotating member is a first carrier 28 that is pivotally supported.

  Similarly, the second planetary gear set 16 is composed of rotating members such as a second sun gear 30, a second ring gear 32, a plurality of second pinions 34, and a second carrier 38, and the third planetary gear set 18a The third sun gear 40, the third ring gear 42, a plurality of third pinions 44, and a third carrier 48 are included. The fourth planetary gear set 18b includes a fourth sun gear 40, a fourth ring gear 42, and a plurality of second gears. It is composed of rotating members such as a 4-pinion 44 and a fourth carrier 48.

Here, the third planetary gear set 18a and the fourth planetary gear set 18b are connected to each other by rotating members to constitute the planetary gear group 18 of the present invention, and each rotating member of the present invention is configured as follows. It is composed.
That is, the third ring gear 42 is the input member of the present invention, the third carrier 48 and the fourth ring gear 52 are connected to each other to output the member of the present invention, the fourth carrier 58 is the first reaction force member of the present invention, The 3rd sun gear 40 and the 4th sun gear 50 connect, and constitute the 2nd reaction force member of the present invention, respectively.

Next, the connection relationship between the first planetary gear set 14, the second planetary gear set 16, and the planetary gear group 18 including the third planetary gear set 18a and the fourth planetary gear set 18b will be described below.
The first carrier 28 is connected to the input shaft 10 and can be connected to the second sun gear 30 by the first clutch 60. The first sun gear 20 can be connected to the second sun gear 30 by the second clutch 64, and The first brake 66 can be fixed to a transmission case (stationary portion) 68.

The first ring gear 22 can be connected to the third ring gear 42 which is the input member of the present invention by the third clutch 70.
The second carrier 38 can be connected to the fourth carrier 58 that is the first reaction force member of the present invention by the fifth clutch 76.

The second ring gear 32 is connected to the third sun gear 40 and the fourth sun gear 50, which are the second reaction force members of the present invention, and can be connected to the second sun gear 30 by the fourth clutch 72. When the fourth clutch 72 is engaged, the second planetary gear set 16 is integrated (a state in which the rotating member of the second planetary gear set 16 can rotate integrally without relative rotation).
The fourth carrier 58 can be fixed to the case 68 by the second brake 74.
The third carrier 48 and the fourth ring gear 52 which are output members of the present invention are connected to the output shaft 12.

  When torque is transmitted from the third ring gear 42 which is the input member of the present invention to the third carrier 48 and the fourth ring gear 52 which are the output members of the present invention, the fourth carrier 58 of the first reaction force member. Is coupled to the second carrier 38 or fixed to the case 68, the first reaction force acting on the fourth carrier 58 of the first reaction force member is the third reaction gear 4 and the fourth sun gear of the second reaction force member described later. It becomes larger than the second reaction force acting on 50. On the other hand, when the fourth carrier 58 is not connected to the second carrier 38 or not fixed to the case 68 and the reaction force is applied to the third sun gear 40 and the fourth sun gear 50 that are the second reaction force members, The second reaction force torque is smaller than the first reaction force acting on the first reaction force member.

Further, as will be described later, the third clutch 70 continues to be engaged at each forward shift position, and is released only during reverse travel. This is to prevent the first sun gear 20 from rotating at a high speed when traveling backward while the first ring gear 22 and the third ring gear 42 of the input member are connected.
Therefore, the third clutch 70 does not need to be switched between engagement and disengagement at the time of shifting during traveling, and does not affect so-called shift shocks.
Therefore, the third clutch 70 can transmit power on the conical friction surface to easily reduce the pressing force required for transmitting torque.
Note that the third clutch 70 may be always engaged by the tension of the spring during forward travel, and may be released by a hydraulic actuator or the like only during reverse travel.

Next, the operation of the planetary gear train of the embodiment shown in FIG. 1 will be described with reference to the operation table shown in FIG.
In the following description, the clutch and the brake are referred to as a friction element.
In the operation table of FIG. 2, friction elements such as clutches and brakes are assigned to the horizontal columns, and C-1 represents the first clutch 60, B-1 represents the first brake 66, and so on. The relationship between these symbols and the symbols of the respective friction elements is shown in FIG.

In the vertical column of the operation table, “D range” and “R range” among the ranges such as “P”, “R”, “N”, “D”, and “L” of an operation lever (not shown) are listed. The reverse range is assigned to the R range from the second forward speed (1st) to the eleventh speed (11th).
In the operation table of FIG. 2, “o” represents the fastening of each fastening element, and the blank represents the release of each fastening element. In addition, the circles in parentheses indicate that they are fastened but are not involved in power transmission.

  Here, the ratio of the number of teeth of each planetary gear set related to the gear ratio is expressed by α, the ratio (Zs / Zr) of the number of teeth (Zs) of the sun gear to the number of teeth (Zr) of the ring gear, and the first planetary gear. The set 14 is described as α1, the second planetary gear set 16 as α2, the third planetary gear set 18a as α3, and the fourth planetary gear set 18b as α4.

Here, for the calculation of each gear ratio, the respective gear ratios are exemplified for α1 of 0.39, α2 of 0.34, α3 of 0.56, and α4 of 0.56.
The gear ratio is represented by the ratio of the rotational speed of the input shaft 10 and the rotational speed of the output shaft 12 (the rotational speed of the input shaft 10 / the rotational speed of the output shaft 12).
Further, in order to simplify the calculation, a formula for calculating the gear ratio is expressed by assuming that the value of 1 / α2−α4−α3 (1 + α4) is A. In the above gear ratio, A is 1.508.

In advance, the third clutch 70 (C-3) is always engaged as described above.
First, in the forward first speed (1st), in addition to the engagement of the third clutch 70, the first clutch 60 (C-1), the fifth clutch 76 (C-5), and the second brake 74 (B- This is done by the conclusion of 2).
The subsequent engagement of the second brake 74 is maintained up to the fifth speed.
As shown in the operation table of FIG. 2, the third clutch 70 is engaged at the first speed, but is not involved in power transmission.
The speed ratio of the first speed is 1 / α2 · α4, and the gear ratio set to the above value is 5.252.

Next, the shift to the forward second speed (2nd) is performed by releasing the engagement of the first clutch 60 at the first speed and engaging the second clutch 64 (C-2).
In the subsequent shift, the engagement of the second clutch 64 is maintained up to the seventh speed.
The speed ratio of the second speed is (1 + α3) / (1 + α1) + {α1 / α2 (1 + α1) + α3 / (1 + α1)} / α4, and the gear ratio set to the above value is 3.315. .

Next, the shift to the third speed (3rd) is performed by releasing the engagement of the fifth clutch 76 at the second speed and engaging the first clutch 60.
The speed ratio of the third speed is 1 + α3 + α3 / α4, which is 2.560 in the above-described gear ratio.

Next, the shift to the fourth speed (4th) is performed by releasing the engagement of the first clutch 60 at the third speed and engaging the first brake 66 (B-1).
The gear ratio is (1 + α3) / (1 + α1) + α3 / {α4 (1 + α1)}.
In the above-mentioned tooth number ratio, it is 1.842.

Next, the shift to the fifth speed (5th) is performed by releasing the engagement of the first brake 66 at the fourth speed and engaging the fourth clutch 72.
The speed ratio of the fifth speed is (1 + α3) / (1 + α1) + {α3 / (1 + α1) −α1 / (1 + α1)} / α4. In the above-mentioned tooth number ratio, it is 1.341.

Next, the shift to the sixth speed (6th) is performed by releasing the engagement of the second brake 74 up to the fifth speed and engaging the first brake 66.
The speed ratio of the sixth speed is (1 + α3) / (1 + α1). In the above-mentioned tooth number ratio, it is 1.122.

Next, the shift to the seventh speed (7th) is performed by releasing the engagement of the first brake 66 and engaging the first clutch 60 at the sixth speed.
As a result, the entire planetary gear train is integrated, and the input shaft 10 and the output shaft 12 are directly connected, so the gear ratio of the sixth speed is 1 regardless of the gear ratio.

Next, the shift to the eighth speed (8th) is performed by releasing the engagement of the second clutch 64 at the seventh speed and engaging the first brake 66.
The speed ratio of the eighth speed is (1 + α3) / (1 + α1 + α3). In the above-mentioned tooth number ratio, the speed increase is 0.800 (overdrive).

Next, the shift to the ninth speed (9th) is performed by releasing the engagement of the first clutch 60 at the eighth speed and engaging the fifth clutch 76.
As a result, the gear ratio becomes 1 / (1 + α1), and the above-mentioned gear ratio becomes 0.719.

Next, the shift to the 10th speed (10th) is performed by releasing the engagement of the fourth clutch 72 at the 9th speed and engaging the first clutch 60 again.
As a result, the gear ratio becomes (A / α1) / (A / α1 + 1 / α2-α4), and the above-mentioned gear ratio becomes 0.619.

Next, the shift to the 11th speed (11th) is performed by releasing the engagement of the first clutch 60 and engaging the second clutch 64 at the 10th speed.
As a result, the gear ratio becomes {(A / (1 + α1)) / (1 / α2−α4), and the speed increase is 0.455 in the above-described gear ratio.

Next, the reverse shift of the R range is performed by releasing the engagement of the third clutch 70 and engaging the first clutch 60, the fourth clutch 72, and the second brake 74.
As a result, the gear ratio becomes -1 / α4, and in the above-mentioned gear ratio, it becomes -2.941.

The following is a list of the forward gear ratios described above. The value on the left is the gear ratio, and the value in the right parenthesis is the ratio between the gear ratio and the gear ratio one step higher (interstage ratio).
Second Speed 5.252 (1.584)
2nd speed 3.315 (1.295)
3rd speed 2.560 (1.390)
4th speed 1.842 (1.374)
5th speed 1.341 (1.195)
6th speed 1.122 (1.122)
7th speed 1.000 (1.250)
8th speed 0.800 (1.112)
9th speed 0.719 (1.163)
10th speed 0.619 (1.359)
11th speed 0.455

From this, it can be seen that the ratio between each stage from the fifth speed to the tenth speed is slightly small. Therefore, when the speed is shifted by one step, the ratio is changed as follows.
5th-7th speed 1.341
6th speed-8th speed 1.403
7th speed-9th speed 1.293

These interstage ratios of about 1.3 to 1.4 can be said to be preferable values for the transmission ratio for automobiles.
In this case, in normal acceleration, control is performed to jump over the sixth speed and the eighth speed. For example, when shifting to a constant speed when accelerating to the seventh speed, the eighth speed or the eighth speed is determined according to the traveling conditions. Since it is conceivable to select one of the 9th gears to change gears, the skipped gear is never wasted.
Rather, since the inter-step ratio from the fifth speed to the tenth speed, which is a frequently used region in steady running, is small, it is possible to travel by selecting the most preferable gear ratio according to the running conditions of steady running. It becomes possible.

Also, instead of using all 11 stages, it is also possible to select from these 11 stages and use it with a reduced number of shift stages such as from the first speed to the tenth speed or from the second speed to the eleventh speed. it can.
In this case, it is preferable to change the gear ratio of each planetary gear set 14, 16, 18a, 18b to a value different from that in the above example so that the gear ratio is suitable for the number of gears.

As described above, the multi-stage planetary gear train according to the first embodiment of the present invention can obtain a preferable gear ratio for an automobile having 11 forward speeds and 1 reverse speed. In addition, since the gear train is configured as described above, the four planetary gears 14, 16, 18a, and 18b can all use a single pinion type that has a simple structure, is light in weight, and has high power transmission efficiency. The number of friction elements that are always free of rotation is 3, which is two less than the conventional example, so drag resistance (drag torque) of the idle friction elements should be reduced accordingly. Can do.
As a result, it is possible to obtain a multi-speed automatic transmission such as a forward 10-speed or 11-speed forward that is low in manufacturing cost, excellent in fuel efficiency, and generates little heat.

FIG. 3 shows a skeleton diagram of a multi-stage planetary gear train according to the second embodiment of the present invention.
Here, the description will focus on parts that are different from the first embodiment, the same reference numerals are given to parts that are substantially the same as those of the first embodiment, and descriptions thereof are omitted.

The difference between the second embodiment and the first embodiment is that the connection relationship between the rotating members of the third planetary gear set 18a and the fourth planetary gear set 18b which are the planetary gear groups of the present invention is different.
That is, the fourth sun gear 50 is the input member of the present invention, the third ring gear 42 and the fourth carrier 58 are connected, and the output member of the present invention is, as will be described later, the third carrier 48 and the fourth ring gear 52. The first reaction force member of the present invention can be connected, and the third sun gear 40 constitutes the second reaction force member of the present invention.

  Second, the third clutch 70 in the first embodiment is omitted, and the first ring gear 22 and the fourth sun gear 50 that is the input member are always connected. Instead, the third carrier 48 and the fourth ring gear 52 are connected. The third carrier 48 can be fixed to the case 68 independently when both are not connected.

That is, the sleeve 78 is provided in the third carrier 48. When the sleeve 78 is moved to the right side in the drawing by an actuator (not shown), the third carrier 48 is engaged with the dog clutch 52a provided in the fourth ring gear 52, and the third carrier 48 is in contact with the fourth ring gear 52. When connected and moved to the left, the third carrier 48 is fixed to the case 68 by meshing with a dog clutch 68 a provided on the case 68.
FIG. 3 is drawn with the sleeve 78 moved to the right and connected to the fourth ring gear 52.

The third difference is that a fourth clutch that integrates the second planetary gear set 16 is provided between the second sun gear 30 and the second carrier 38.
Other than this, the connection relationship among the rotating members of the planetary gear group and the rotating members of the input shaft 10, the output shaft 12, the first planetary gear set 14, and the second planetary gear set 16 is the same as that of the first embodiment. .

Next, the operation of the second embodiment will be described.
The operation table of Example 2 is shown in FIG. The difference from the operation table of the first embodiment shown in FIG. 2 is that the third clutch 70 and the sleeve 78 (S) in the first embodiment are switched as described above.
That is, in forward movement, the sleeve 78 moves to the right side to connect the third carrier 48 to the fourth ring gear 52, and in reverse movement, it moves to the left side to fix the third carrier 48 to the case 68. Yes.

  The movement of the sleeve 78 for forward / reverse switching is performed in a state where the vehicle is stopped or moving at a slow speed. Before the sleeve 78 is moved, the second brake 74 is engaged and the fourth ring gear 52 is rotated. By performing the operation after stopping, the meshing can be performed in a state where there is no difference in rotational speed between the sleeve 78 and the dog clutches 68a and 52a.

Since the operation of the friction element at other speed changes is the same as that of the first embodiment, detailed description thereof is omitted.
Although description of the calculation of each gear ratio is omitted, in the second embodiment, similarly to the first embodiment, it is possible to perform the 11th forward and reverse shifts.

Similarly to the first embodiment, the multi-stage planetary gear train according to the second embodiment of the present invention has 11 forward and 1 reverse gear ratios preferable for an automobile, and the gear train is configured as described above. The planetary gears 14, 16, 18 a, and 18 b are all simple structure, light weight, high power transmission efficiency and single pinion type, and there are 6 friction elements. Since the number of friction elements is three, which is two less than that of the conventional example, drag resistance of the friction elements that are idle can be reduced.
As a result, it is possible to obtain a forward 10-speed or 11-speed automatic transmission with low manufacturing cost, excellent fuel efficiency, and low heat generation.

As described above, the multi-stage planetary gear train according to each embodiment of the present invention provides an automatic transmission that has a forward 11-speed ratio that is preferable for an automobile, is low in manufacturing cost, has excellent fuel efficiency, and generates little heat. It becomes possible to obtain.
This is because the connection relationship between the first planetary gear set 14, the second planetary gear set 16, and the planetary gear group 18 is as described above, and the planetary gear group 18 is not limited to the illustrated combination, In general, it may be a planetary gear group that can obtain a forward three-speed gear ratio.

In each of the above embodiments, the torque converter 2 is provided between the engine 1 and the input shaft 10, but it goes without saying that a fluid coupling or a friction clutch may be used instead.
In addition, as is generally done with an automatic transmission, it is also possible to provide a one-way clutch in parallel with a friction element that is fastened at a low speed, thereby facilitating shift control in shifting to a higher gear.

  Since it is possible to obtain an automatic transmission with 10 or 11 forward gear ratios, low manufacturing costs, excellent fuel efficiency, and low heat generation, from small passenger cars where fuel efficiency is important to medium-sized commercial vehicles, etc. Can be widely applied.

It is the skeleton figure which showed the multi-stage planetary gear train of this invention. Example 1 It is a figure which shows the operation | movement table | surface of the multistage speed planetary gear train of Example 1. FIG. It is the skeleton figure which showed the multi-stage planetary gear train of this invention. (Example 2) It is a figure which shows the operation | movement table | surface of the multistage speed planetary gear train of Example 2. FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine 2 Torque converter 10 Input shaft 12 Output shaft 14 1st planetary gear set 16 2nd planetary gear set 18 Planetary gear set group 18a 3rd planetary gear set 18b 4th planetary gear set 20 1st sun gear 22 1st ring gear 24 1st 1 pinion 28 1st carrier 30 2nd sun gear 32 2nd ring gear 34 2nd pinion,
38 second carrier 40 third sun gear 42 third ring gear 44 third pinion 48 third carrier 50 fourth sun gear 52 fourth ring gear 54 fourth pinion 58 fifth carrier 60 first clutch 64 second clutch 66 first brake 68 case 70 Third clutch 72 Fourth clutch 74 Second brake 76 Fifth clutch 72
78 sleeve

Claims (4)

An input shaft;
An output shaft;
A first planetary gear that is coaxially arranged with the input shaft and includes a first sun gear, a first ring gear, a first ring gear, and a first pinion that meshes with the first sun gear, and a first carrier that rotatably supports the first pinion. A gear set,
A second carrier that is disposed between the input shaft and the output shaft and rotatably supports the second sun gear, the second ring gear, the second ring gear, and the second sun gear. A second planetary gear set provided;
A first reaction force member that is disposed coaxially with the output shaft, receives a reaction torque when transmitting torque from the input member to the output member, and from the input member to the output member. A planetary gear group including a second reaction force member that receives a reaction force torque smaller than a reaction force torque that the first reaction force member receives when transmitting torque to
The input shaft is connected to the first carrier and connectable to the second sun gear,
The first ring gear is connected to or connectable to the input member;
The first sun gear can be coupled to the second sun gear and can be fixed to a stationary portion;
The first reaction force member can be connected to the second carrier and fixed to a stationary part,
The second reaction force member is connected to the second ring gear;
The output member is connected to the output shaft;
A multi-speed planetary gear train comprising a clutch capable of integrating the second planetary gear set. The planetary gear group includes a third sun gear, a third ring gear, a third ring gear, a third pinion meshing with the third sun gear, and a third planetary gear set having a third carrier that rotatably supports the third pinion. , A fourth sun gear, a fourth ring gear, a fourth ring gear, a fourth pinion meshing with the fourth sun gear, and a fourth planetary gear set having a fourth carrier that rotatably supports the fourth pinion,
The third ring gear constitutes the input member, the third carrier and the fourth ring gear are connected to each other to constitute the output member, the fourth carrier constitutes the first reaction force member, The multi-stage planetary gear train according to claim 1, wherein the third sun gear and the fourth sun gear are connected to each other to form the second reaction force member. The planetary gear group includes a third sun gear, a third ring gear, a third ring gear, a third pinion meshing with the third sun gear, and a third planetary gear set having a third carrier that rotatably supports the third pinion. , A fourth sun gear, a fourth ring gear, a fourth ring gear, a fourth pinion meshing with the fourth sun gear, and a fourth planetary gear set having a fourth carrier that rotatably supports the fourth pinion.
The fourth sun gear constitutes the input member, the third ring gear and the fourth carrier are connected to each other to constitute the output member, and the third carrier and the fourth ring gear can be connected to each other. The multi-speed planetary gear train according to claim 1, wherein the first reaction force member is configured, and the third sun gear is the second reaction force member.   Means capable of connecting the third carrier and the fourth ring gear, and means for fixing the third carrier to a stationary part when the third carrier and the fourth ring gear are not connected are provided. The multi-speed planetary gear train according to claim 3.

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