JP2009293762A - Multistage shift planetary gear train - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve fuel economy by enhancing power transmitting efficiency in a multistage shift planetary gear train for forward 8-speeds. <P>SOLUTION: A multistage shift planetary gear train includes an input shaft 10, an output shaft 12, a first and second planetary gear sets 14, 16, a first reduction planetary gear set 18 including an input member 40, an output member 48 and a holding member 42 and a second reduction planetary gear 19. The input shaft 10 is connected to a first carrier 28 and can be connected to the input member 40 and second ring gear 32 integrally connected to each other. The output shaft 12 can be connected to a second carrier 38 and to the output member 48 through the second reduction planetary gear 19. A first sun gear 20 and the holding member 42 are integrally connected to each other and can be held to a stationary part 64. A first ring gear 22 is integrally connected with a second sun gear 30. A clutch 66 for integrating the second planetary gear set 16 is included in the planetary gear train. <P>COPYRIGHT: (C)2010,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 for use in automatic transmissions for vehicles, gears capable of multi-speed shifting with eight forward speeds have been put into practical use, 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 set of double-pinion type planetary gears and a planetary gear group called Ravigneaux type, and a multi-speed planetary gear train consisting of six friction elements. The gear train obtains a gear ratio of eight forward speeds by switching so that two of the six friction elements are always engaged. (See Patent Document 1).

However, since the above conventional planetary gear train requires two sets of double pinion type planetary gears in order to obtain a preferable gear ratio for a forward eight-speed automobile, it is manufactured in comparison with a so-called single pinion type planetary gear. While there is a problem that the cost increases and the gear meshing is large, the power transmission efficiency is poor. On the other hand, four of the six friction elements are always idle, so that the drag of the friction element that is idle is dragged. Combined with the problem that the torque is increased and the power transmission efficiency is poor, there is a problem that the fuel consumption of the automobile is bad and the heat generation is large.
Japanese Patent No. 3777929

The problem to be solved is that two sets of double pinion type planetary gears with poor power transmission efficiency are required, and four friction elements always rotate freely, resulting in poor fuel consumption and high heat generation. .
The object of the present invention is to make the double pinion type planetary gears to rotate at one set or less, i.e., one set or less, and the rest using single pinion type planetary gears, while ensuring a preferable gear ratio for a vehicle such as an automobile, etc. It is an object of the present invention to provide a multi-speed planetary gear train capable of reducing fuel consumption and heat generation by reducing the number of friction elements.

  The multi-stage planetary gear train of the present invention includes an input shaft, an output shaft, a first pinion that is arranged coaxially with the input shaft, and meshes with the first sun gear, the first ring gear, the first ring gear, and the first sun gear. A first planetary gear set having a first carrier that rotatably supports one pinion, a second sun gear, a second ring gear, a second pinion meshed with the second ring gear and the second sun gear, and rotating the second pinion A second planetary gear set having a second carrier that freely supports the shaft, a first reduction planetary gear set having an input member, an output member, and a fixed member, and a second reduction gear, and the input shaft being a first carrier The input member and the second ring gear that are connected and integrally connected to each other can be connected, and the output shaft can be connected to the second carrier and the output member and the second reduction gear. The first sun gear and the fixed member can be connected together and fixed to the stationary part, the first ring gear can be connected to the second sun gear and the second planetary gear. The clutch which united the set 16 was provided.

  Since the multi-stage planetary gear train of the present invention is configured as described above, the double pinion type planetary gear is suppressed to one set or less, and it is preferable for a transmission for an automobile or the like using a single pinion type having high power transmission efficiency. While obtaining the gear ratio, the number of friction elements that are always idle can be reduced by two as compared with the conventional example, so that improvement in fuel consumption and heat generation can be expected.

  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).

The first planetary gear set 14, the second planetary gear set 16, the third planetary gear set 18 and the fourth planetary gear set 19 constituting the gear train are all generally called single pinion types, Have 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. The fourth planetary gear set 19 includes a fourth sun gear 50, a fourth ring gear 52, and a plurality of first gears 19. The fourth planetary gear set 19 includes a third sun gear 40, a third ring gear 42, a plurality of third pinions 44, and a third carrier 48. It is composed of rotating members such as a 4-pinion 54 and a fourth carrier 58.

Next, the connection relationship between the first planetary gear set 14, the second planetary gear set 16, the third planetary gear set 18, and the fourth planetary gear set 19 will be described below.
The input shaft 10 is connected to the first carrier 28 and is connectable to the second ring gear 32 and the third sun gear 40 that are connected to each other via the first clutch 60.

The first sun gear 20 is integrally connected to the third ring gear 42 and can be fixed to a transmission case (stationary portion) 64 by a first brake 62.
The first ring gear 22 is connected to the second sun gear 30 and can be connected to the second ring gear 32 and the third sun gear 40 that are connected to each other via the second clutch 66.

When the second clutch 66 is engaged, the second sun gear 30 and the second ring gear 32 are connected, so that the second planetary gear set 16 is integrated (the rotating members of the second planetary gear set 16 are rotated together). Possible).
The third carrier 48 is connected to the fourth sun gear 50.
The fourth ring gear 52 can be fixed to the transmission case 64 by the second brake 70.
The output shaft 12 is connected to the fourth carrier 58 and can be connected to the second carrier 38 via the third clutch 68.

Here, the third planetary gear set 18 constitutes the first reduction planetary gear set of the present invention, the third sun gear 40 is the input member of the present invention, the third carrier 48 is the output member of the present invention, and the third ring gear. 42 respectively constitutes the fixed members of the present invention.
The fourth planetary gear set 19 constitutes the second reduction gear of the present invention.
That is, when the fourth ring gear 52 is fixed to the case 64 by fastening the second brake 70, the input to the fourth sun gear 50 can be decelerated and output to the output shaft 12.

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, C-1 represents the first clutch 60, B-1 represents the first brake 62, 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 taken up. The R range from forward 1st speed (1st) to 8th speed (8th) is assigned to each reverse speed.
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.

  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 group 14 will be described as α1, the second planetary gear group 16 as α2, the third planetary gear group 18 as α3, and the fourth planetary gear group 19 as α4.

Here, the respective gear ratios used for the calculation of the respective gear ratios are exemplified for α1 of 0.60, α2 of 0.60, α3 of 0.60, and α4 of 0.55.
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).
In order to simplify the calculation formula, α1 · α2 (1 + α3) / (1 + α2) will be described as A.
In the above-mentioned tooth number ratio, A is 0.360.

First, as shown in the operation table shown in FIG. 2, the first forward speed (1st) is driven by the first clutch 60 (C-1), the first brake 62 (B-1), and the second brake 70 ( It is performed by the conclusion of B-2).
In the subsequent shift, the engagement of the second brake 70 is maintained up to the fifth speed.
The speed ratio of the first speed is (1 + α3) (1 + α4) / (α3 · α4), and the gear ratio set to the above value is 7.515.

Next, the shift to the second speed (2nd) is performed by releasing the engagement of the first clutch 60 at the first speed and engaging the second clutch 66 (C-2).
The speed ratio of the second speed is (1 + α3) (1 + α4) / {α3 · α4 (1 + α1)}, which is 4.697 in the above-described gear ratio.

Next, the shift to the third speed (3rd) is performed by releasing the engagement of the first brake 62 at the second speed and engaging the first clutch 60 again.
The gear ratio is (1 + α4) / α4.
In the above-mentioned tooth number ratio, it is 2.818.

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 third clutch 68 (C-3).
The speed ratio of the fourth speed is {α4 (1 + α1) + α1 (1 + α3)} / {α4 (1 + α1)}. In the above-mentioned tooth number ratio, it is 2.091.

Next, the shift to the fifth speed (5th) is performed by releasing the engagement of the second clutch 66 at the fourth speed and engaging the first clutch 60 again.
The speed ratio of the fifth speed is {α4 (1 + A) + A} / {α4 (1 + A)}. In the above-mentioned tooth number ratio, it is 1.481.

Next, the shift to the sixth speed (6th) is performed by releasing the engagement of the second brake 70 up to the fifth speed and engaging the second clutch 66.
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 seventh speed (7th) is performed by releasing the engagement of the second clutch 66 at the sixth speed and engaging the first brake 62.
As a result, the gear ratio becomes (1 + α3) / (1 + α3 + A), and the speed increase is 0.816 in the above-described gear ratio.

Next, the shift to the eighth speed (8th) is performed by releasing the engagement of the first clutch 60 at the seventh speed and engaging the second clutch 66 again.
As a result, the gear ratio becomes 1 / (1 + α1), and the gear ratio is an increase of 0.625.

Next, the reverse shift of the R range is performed by engaging the third clutch 68, the first brake 62, and the second brake 70.
As a result, the gear ratio becomes (1 + α2) / {α2 (1 + α1)} − (1 + α3) (1 + α4) / {α2 · α3 · α4 (1 + α1)}, and in the above-described gear ratio, −6.162. is there.

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).
1st speed 7.515 (1.600)
Second speed 4.697 (1.667)
3rd speed 2.818 (1.348)
4th speed 2.091 (1.412)
5th speed 1.481 (1.481)
6th speed 1.000 (1.225)
7th speed 0.816 (1.306)
8th speed 0.625

  From this, it can be seen that a gear ratio of 8 steps, which is a preferable gear ratio for a large truck or bus having a heavy vehicle gross weight, can be obtained.

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 the automobile in the 8 forward speeds and the 1 reverse speed, and the gear train is configured as described above. The gear sets 14, 16, 18, and 19 can use a single pinion type that has a simple structure, is light in weight, and has high power transmission efficiency without using a double pinion type at all. The number is 2 and 2 less than the conventional example, so that drag resistance (drag torque) of the idle friction element can be reduced.
As a result, it is possible to obtain an eight-speed automatic transmission with excellent fuel efficiency and low heat generation.

FIG. 3 shows a skeleton diagram of the 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 double pinion type is used while being limited to only one set, and the third planetary gear set 18 constituting the first reduction planetary gear of the present invention is a double pinion type. It is that you are.
That is, the third planetary gear set 18 includes a third sun gear 40, a third ring gear 42, an outer pinion 44 engaged with the third ring gear 42, an inner pinion 46 engaged with the outer pinion 44 and the third sun gear 40, and the outer The third carrier 48 is configured to rotatably support the pinion 44 and the inner pinion 46. The third carrier 48 is the input member of the present invention, the third sun gear 40 is the fixed member of the present invention, and the third ring gear 42 is the main. Configure the output member of the invention.

Other configurations and connection relationships between the members are the same as those in the first embodiment, and thus detailed description thereof is omitted.
Next, the operation of the planetary gear train of the second embodiment shown in FIG. 3 is the same as that of the first embodiment shown in FIG.

Since the third planetary gear set 18 is of a double pinion type, the calculation formula for each gear ratio changes as follows.
A: α4 {α3 (1 + α2) + α1 · α2}
First gear: (1 + α4) / {α4 (1-α3)}
Second speed: (1 + α4) / {α4 (1 + α1) (1-α3)}
3rd speed: (1 + α4) / α4
Fourth gear: {α3 · α4 (1-α1) + α1} / {α3 · α4 (1 + α1)}
5th speed: (α1 ・ α2 + A) / A
6th speed: 1
7th speed: (1 + α2) / (1 + α2 + α1 ・ α2)
8th speed: 1 / (1 + α1)
Reverse: (1 + α2) / {α2 (1 + α1)}-(1 + α4) / {α2 · α4 (1 + α1) (1-α3)}

When the ratio of the number of teeth of the first to fourth planetary gear sets 14, 16, 18, 19 is α10 is 0.410, α2 is 0.626, α3 is 0.460, and α4 is 0.545. The gear ratio is as follows. 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).
A 0.548
1st speed 5.250 (1.410)
2nd speed 3.723 (1.313)
3rd speed 2.835 (1.313)
Fourth speed 2.160 (1.471)
5th speed 1.469 (1.469)
6th speed 1.000 (1.158)
7th speed 0.864 (1.219)
8th speed 0.709
Reverse -4.105

  From this, it can be seen that a gear ratio of eight steps, which is a preferable step ratio for a passenger car, can be obtained.

As described above, the multi-stage planetary gear train according to the second embodiment of the present invention also provides a preferable gear ratio for the automobile in the forward 8th and the reverse 1st, and the gear train is configured as described above. Of the gear sets 14, 16, 18 and 19, the double pinion type is limited to only one set. Others can use a single pinion type with a simple structure, light weight and high power transmission efficiency, and is always free to rotate. Since the number of friction elements is two, which is two fewer 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 an eight-speed automatic transmission with excellent fuel efficiency and low heat generation.

FIG. 4 shows a skeleton diagram of a multi-stage planetary gear train according to the third embodiment of the present invention. In FIG. 4, the input shaft 10 side shows the upper half of the input shaft 10, and the output shaft 12 side shows the lower half of the output shaft 12.
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 third embodiment and the first embodiment is that the output pin 12 is parallel to the input shaft 10 in that the double pinion type is used while being limited to only one set and that two sets of planetary gears are arranged in a radial direction. Accordingly, the fourth carrier 58 and the second carrier 38 and the output shaft 12 are connected by the connecting gear pair 72, and the third planetary gear set is the same as in the second embodiment. 18 is a double pinion type, and the second planetary gear set 16 is disposed radially outside the first planetary gear set 14 so that the first ring gear 22 and the second sun gear 30 are integrated. .

Next, the operation of the planetary gear train of the third embodiment shown in FIG. 4 is the same as that of the first embodiment shown in FIG.
As described above, since the coupling gear pair 72 is provided between the fourth carrier 58 and the output shaft 12, the gear ratio of the coupling gear pair 72 is related to the calculation of the transmission ratio. Since the calculation formula is simply multiplied by the gear ratio of the coupling gear pair 72, detailed description thereof is omitted.

Therefore, the third embodiment shown in FIG. 4 can also obtain a gear ratio of eight steps, which is a preferable step ratio as a gear ratio for passenger cars, as in the second embodiment.
Further, since the second planetary gear set 16 is arranged on the radially outer side of the first planetary gear set 14, there is a merit of shortening the axial length of the gear train. For this reason, it can be said that Example 3 is suitable for a passenger car such as a so-called front-wheel drive of a horizontal engine.

As described above, the multi-stage planetary gear train according to the third embodiment of the present invention also provides a preferable gear ratio for the automobile in the forward 8th and the reverse 1st, and the gear train is configured as described above. Of the gear sets 14, 16, 18 and 19, the double pinion type is limited to only one set. Others can use a single pinion type with a simple structure, light weight and high power transmission efficiency, and is always free to rotate. Since the number of friction elements is two, which is two less than the conventional example, the drag resistance of the friction elements that are idle can be reduced.
As a result, it is possible to obtain an eight-speed automatic transmission with excellent fuel efficiency and low heat generation.

FIG. 5 shows a skeleton diagram of a multi-stage planetary gear train according to the fourth embodiment of the present invention. 5 is the same as that in the third embodiment.
Here, the description will focus on the parts different from the first and third embodiments, the same reference numerals are given to the substantially same parts, and the description thereof is omitted.

The fourth embodiment is the same as the third embodiment in that the output shaft 12 is provided in parallel with the input shaft 10, but the second reduction gear of the present invention is not a planetary gear set but a reduction gear pair 74. Is different from Example 1 and Example 3.
The first planetary gear set 14, the second planetary gear set 16, and the third planetary gear set 18 are arranged in series on the input shaft 10 in the same manner as in the first embodiment, but the third planetary gear set 18 is double. The pinion type is the same as in the second embodiment.

The third clutch 68 is disposed on the output shaft 12 side, and the fourth clutch 76 provided between the reduction gear pair 74 and the output shaft 12 is also disposed on the output shaft 12 side.
The fourth clutch 76 performs the same function as the second brake 70 in the first embodiment.
That is, by engaging the second brake 70, the third ring gear 42 that is an output member and the output shaft 12 are connected by the reduction gear pair 74.

  Further, the relationship of the rotating members of the third planetary gear set 18 which is the first reduction planetary gear set of the present invention is different from those in the second and third embodiments. That is, the third sun gear 40 and the third carrier 48 are interchanged, and the third sun gear 40 constitutes the input member of the present invention, and the third carrier 48 constitutes the fixed member of the present invention.

Next, the operation of the planetary gear train of the fourth embodiment shown in FIG. 5 will be described with reference to the operation table shown in FIG.
As can be seen by comparing FIG. 6 with the operation table of the first embodiment shown in FIG. 2, the second brake 70 (B-2) in the first embodiment and the fourth clutch 76 (C-4) in the fourth embodiment are switched. It ’s just that.
For this reason, since shifting of engagement and release of each friction element can be performed according to the operation table shown in FIG.

Next, regarding the gear ratio, the second reduction gear of the present invention has been replaced with the reduction gear pair 74, and the third sun gear 40 and the third carrier 48 of the third planetary gear set 18 have been replaced. The gear ratio calculation formula changes as follows.
Here, the ratio of the number of teeth of the coupling gear pair 72 and the reduction gear pair 74 (the number of teeth on the input shaft 10 side / the number of teeth on the output shaft 12 side) is i1 for the coupling gear pair 72, and the reduction gear pair 74. In this case, i2 is defined and described.

In order to simplify the calculation formula, 1 / (1 + α1) −α1 (1-α3) / {α3 (1 + α1)} is B, α3 · α2 / {α3 + α2 · α3 (1 + α1) + α1 · α2 (1-α3)} Is represented as C.
1st speed: i2 / α3
Second speed: i2 / {α3 (1 + α1)}
3rd speed: i2
Fourth speed: α1 · i2 / {(1 + α1) (1-α3)} + i1 · B
5th speed: α1 · i2 · C / α3 + i1 · C (1 + α2) / α2
6th speed: i1
7th speed: i1 (1 + α2) / {1 + α2 (1 + α1)}
8th speed: i1 / (1 + α1)
Reverse: i1 (1 + α2) / {α2 (1 + α1)} − i2 / {α2 · α3 (1 + α1)}

In addition, the gear ratio of the first to fourth planetary gear sets 14, 16, 18, and 19 is as follows: α1 is 0.430, α2 is 0.626, α3 is 0.490, i1 is 0.90, and i2 is In the case of 1.98, the gear ratio is as follows. 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).
C 0.386
B 0.288
1st speed 4.041 (1.430)
Second speed 2.826 (1.427)
3rd speed 1.980 (1.307)
4th speed 1.515 (1.292)
5th speed 1.173 (1.303)
6th speed 0.900 (1.166)
7th speed 0.772 (1.227)
8th speed 0.629
Reverse -2.879

From this, it can be seen that a gear ratio of eight steps, which is a preferable step ratio for a passenger car, can be obtained.
Further, similarly to the third embodiment, since the second planetary gear set 16 is arranged radially outside the first planetary gear set 14, there is an advantage of reducing the axial length of the gear train. For this reason, the fourth embodiment is also suitable for a passenger car such as a so-called engine-side-mounted front wheel drive.

As described above, the multi-stage planetary gear train according to the fourth embodiment of the present invention also has a preferable gear ratio for the automobile in the forward 8 speed and the reverse 1 speed, and the gear train is configured as described above. Of the gear sets 14, 16 and 18, the double pinion type is limited to one set, and the other is a single pinion type with a simple structure, light weight and high power transmission efficiency, and is always idle. Since the number of friction elements is two, which is two fewer than the conventional example, drag resistance of the friction elements that are idle can be reduced.
As a result, it is possible to obtain an eight-speed automatic transmission with excellent fuel efficiency and low heat generation.

  As described above, the multi-stage planetary gear train according to each of the embodiments of the present invention can provide an eight-speed forward gear ratio that is preferable for an automobile, and can provide an automatic transmission that is excellent in fuel efficiency and generates less heat. .

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.
The arrangement of each planetary gear set, gear pair, and each friction element can be changed as appropriate according to the layout of the transmission.

  Since it is possible to obtain an automatic transmission that has an eight-speed forward gear ratio, excellent fuel efficiency, and little heat generation, it can be widely applied to small passenger cars, which emphasize fuel efficiency, to large commercial vehicles.

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 the skeleton figure which showed the multi-stage planetary gear train of this invention. (Example 3) It is the skeleton figure which showed the multi-stage planetary gear train of this invention. (Example 4) It is a figure which shows the operation | movement table | surface of the multi-stage speed planetary gear train of Example 4.

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 group 18 3rd planetary gear set 19 4th planetary gear set 20 1st sun gear 22 1st ring gear 24 1st Pinion 28 First carrier 30 Second sun gear 32 Second ring gear 34 Second pinion 38 Second carrier 40 Third sun gear 42 Third ring gear 44 Third pinion, outer pinion 46 Inner pinion 48 Third carrier 50 Fourth sun gear 52 Fourth Ring gear 54 4th pinion 58 4th carrier 60 1st clutch 62 1st brake 64 Case 66 2nd clutch 68 3rd clutch 70 2nd brake 72 Connection gear pair 74 Reduction gear pair 76 4th clutch

Claims (6)

An input shaft;
An output shaft;
A first sun gear, a first ring gear, a first pinion meshing with the first ring gear and the first sun gear, and a first carrier that rotatably supports the first pinion are arranged coaxially with the input shaft. One planetary gear set,
A second planetary gear set having a second sun gear, a second ring gear, a second pinion meshed with the second ring gear and the second sun gear, and a second carrier rotatably supporting the second pinion;
A first reduction planetary gear set having an input member, an output member, and a fixed member;
A second reduction gear,
The input shaft is connectable to the input member and the second ring gear, which are connected to the first carrier and integrally connected to each other,
The output shaft can be connected to the second carrier and can be connected to the output member via the second reduction gear,
The first sun gear and the fixed member are integrally connected and can be fixed to a stationary part,
The first ring gear is integrally connected to the second sun gear,
A multi-speed planetary gear train comprising a clutch that integrates the second planetary gear set.   The first reduction planetary gear set includes a third sun gear, a third ring gear, a third pinion meshed with the third ring gear and the third sun gear, and a third carrier that rotatably supports the third pinion, The multi-stage planetary gear train according to claim 1, wherein the third sun gear constitutes the input member, the third carrier constitutes the output member, and the third ring gear constitutes the fixed member.   The first reduction planetary gear set includes a third sun gear, a third ring gear, an outer pinion meshed with the third ring gear, an inner pinion meshed with the outer pinion and the third sun gear, and the outer pinion and the inner pinion freely rotatable. The third sun gear and the third carrier, one of which constitutes the input member, the other constitutes the fixed member, and the third ring gear constitutes the output member. The multi-speed planetary gear train according to claim 1, wherein   The second reduction gear includes a fourth sun gear, a fourth ring gear, a fourth pinion meshed with the fourth ring gear, and the fourth sun gear, and a planetary gear having a fourth carrier that rotatably supports the fourth pinion. The first sun gear is coupled to the output member, the fourth carrier is coupled to the output shaft, and the fourth carrier is fixable to a stationary part. The multi-speed planetary gear train as set forth in claim 3.   The input shaft and the output shaft are arranged in parallel to each other, the first planetary gear set and the reduction planetary gear set are arranged coaxially with the input shaft, and radially outward of the first planetary gear set. 4. The multi-speed planetary gear train according to claim 1, wherein the second planetary gear set is arranged so that the first ring gear and the second sun gear are integrated. 5.   6. The multi-speed planetary gear train according to claim 5, wherein the output member and the output shaft are connectable via a reduction gear pair constituting the second reduction gear.

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