JP2009063139A - Multistage shift planetary gear train - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To obtain a multistage shift planetary gear train for more than nine forward speed stages, in a transmission having short axial length. <P>SOLUTION: The multistage variable-speed planet gear train is constituted in such a manner that: a first planet gear set 24 is arranged at a first output shaft 12; a second planet gear set 26 is arranged at a second output shaft 14 which is parallel with the first output shaft; an input shaft 10 is connectable to a first sun gear 30 and a first ring gear 32 via a speed-reducing gear set 52, and also connectable to a second ring gear 42 and a second sun gear 40; the first output shaft 12 and the second output shaft 14 are connected to each other via a connecting gear set; the first output shaft 12 and the second output shaft are connected to a first carrier 38 and the second ring gear 42, respectively; the first sun gear 30 and the second sun gear 40 are connected to each other via a second gear set 60, and fixable to a stationary part 64; a second carrier 48 is fixable to the stationary part 64; and the first planet gear set 24 can be integrally formed in a state that the input shaft 10, the first sun gear 30 and the first ring gear 32 are not connected to one another. <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.

A planetary gear train used for a vehicular automatic transmission has been proposed that is capable of multi-speed shifting of eight forward speeds, mainly for improving the fuel efficiency, exhaust characteristics, acceleration performance, and the like of the vehicle.
As a conventional planetary gear train capable of such multi-speed shifting, there is a multi-speed planetary gear train proposed by the present applicant, and these gear trains are shafts suitable for so-called engine-side-mounted vehicles such as front-wheel drive vehicles. In order to obtain a transmission with a short directional length, a planetary gear divided into two output shafts and six friction elements obtain a transmission ratio of eight forward speeds. (See Patent Document 1).

However, since the above conventional planetary gear train has a limit of obtaining a forward gear ratio of eight forward speeds using six friction elements, let's apply a desirable gear ratio according to various running conditions of the vehicle. However, there was a limit, and there was a problem that acceleration performance and fuel consumption could not be further improved.
JP 2005-023987 JP 2005-180665 A

The problem to be solved is that a gear ratio exceeding eight forward speeds cannot be obtained as a gear train applied to a transmission having a short axial length suitable for the engine horizontal type.
An object of the present invention is to provide a multi-speed planetary gear train capable of obtaining a gear ratio exceeding eight forward speeds while keeping the axial length of the planetary gear train short.

  The multi-stage planetary gear train of the present invention includes an input shaft, a first output shaft provided parallel to the input shaft, a second output shaft provided coaxially with the input shaft, or provided parallel to the input shaft and the first output shaft. A first pinion that is arranged coaxially with the first output shaft and that is engaged with the first sun gear, the first ring gear, the first ring gear, and the first sun gear as a rotating member, and the first carrier that rotatably supports the first pinion. The first planetary gear set and the second output shaft are arranged coaxially, and the second sun gear, the second ring gear, the second ring gear, and the second sun gear meshed with the second sun gear as a rotating member can rotate freely. A second planetary gear set comprising a second carrier that supports the shaft is provided, and the input shaft can be connected to the first sun gear and the first ring gear via a speed reduction mechanism, respectively, The first output shaft and the second output shaft can be connected to each other via a first gear set, and the first output shaft can be connected to the first carrier. The second output shaft is connected to the second ring gear, the first sun gear and the second sun gear are connected via the second gear set and can be fixed to the stationary part, and the second carrier can be fixed to the stationary part. In the state where the input shaft and the first sun gear and the first ring gear are not connected, the rotary member of the first planetary gear set can be rotated together.

  Since the multi-speed planetary gear train of the present invention is configured as described above, it is possible to obtain a gear ratio of 10 forward speeds that is a preferable value for a transmission for a vehicle while keeping the axial length of the planetary gear train short. As a result, it is possible to improve the acceleration performance and fuel efficiency of the engine-mounted front-wheel drive vehicle or the like.

  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 an embodiment of the present invention.
In the multi-speed planetary gear train of the embodiment shown in FIG. 1, the input shaft 10 driven from the engine 1 via the torque converter 2 is on the same axis as the output shaft 1a of the engine 1 and in parallel therewith. A first output shaft 12 is disposed, and a second output shaft 14 is disposed coaxially with the input shaft 10. Further, an intermediate output shaft 16 is provided in parallel with these.

The first output shaft 12 and the second output shaft 14 are connected to each other via connecting gears 12 a and 14 a that are integral with each other, and one gear 14 a is also meshed with the input gear 16 a that is integral with the intermediate output shaft 16. These connecting gears 12a and 14a constitute the first gear set of the present invention. The drive gear 16 b integrated with the intermediate output shaft 16 is in mesh with the output gear 18.
The output gear 18 drives the axles 22a and 22b via the differential device 20, and the axles 22a and 22b are connected to left and right wheels (not shown).

A first planetary gear set 24 is coaxially disposed on the first output shaft 12, and a second planetary gear set 26 is coaxially disposed on the second output shaft 14.
Both the first planetary gear set 24 and the second planetary gear set 26 are generally called single pinion types, and each has the same configuration.
That is, the first planetary gear set 24 allows the first sun gear 30, the first ring gear 32, the plurality of first pinions 34 engaged with the first ring gear 32 and the first sun gear 30, and the first pinion 34 to freely rotate. The rotating member is a first carrier 38 that is pivotally supported.

Similarly, the second planetary gear set 26 includes rotating members such as a second sun gear 40, a second ring gear 42, a plurality of second pinions 44, and a second carrier 48.
The input shaft 10 can be connected to an intermediate shaft 54 that is arranged coaxially with the first output shaft via the first clutch 50 and a reduction gear pair (gear 52a, gear 52b). The reduction gear pair constitutes a reduction mechanism of the present invention.
The intermediate shaft 54 can be connected to the first ring gear 32 via the second clutch 56 and can also be connected to the first sun gear 30 via the third clutch 58.

The first sun gear 30 and the second sun gear 40 are connected via a second gear set (gear 60a, gear 60b) and can be fixed to the case (stationary part) 64 by the first brake 62.
The input shaft 10 can be connected to the second carrier 48 via the fourth clutch 66 and to the second sun gear 40 via the fifth clutch 68.
The second carrier 48 can be fixed to the case 64 by the second brake 70. The first output shaft 12 is connected to the first carrier 38, and the second output shaft 14 is connected to the second ring gear 42.

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 operation table of FIG. 2, fastening elements such as clutches and brakes are assigned to the horizontal columns, C-1 represents the first clutch 50, B-1 represents the first brake 62, and so on. The relationship between these symbols and the symbols of the respective fastening elements is shown in FIG.

The vertical column of the operation table is divided into “D range” and “R range” of an operation lever (not shown). The D range is the first forward speed (1st) to the tenth speed (10th), and the R range is reverse ( R-1 and R-2) are assigned.
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, circles enclosed in parentheses indicate that they are not involved in power transmission even if they are fastened.

Here, regarding the calculation of each gear ratio, in the planetary gear set, the ratio (Zs / Zr) of the number of teeth (Zs) of the sun gear to the number of teeth (Zr) of the ring gear is calculated as follows: α1, the second planetary gear set 26 is α2, and in the gear set, the gear ratio of the reduction gear pair (the number of teeth of the gear 52b / the number of teeth of the gear 52a) is i1, and the gear ratio of the first gear set The description will be made assuming that the number of teeth of the connecting gear 12a / the number of teeth of the connecting gear 14a is i2, and the ratio of the number of teeth of the second gear set (the number of teeth of the gear 60b / the number of teeth of the gear 60a) is i3.
The transmission ratio is represented by the ratio of the rotational speed of the input shaft 10 to the rotational speed of the second output shaft 14 (the rotational speed of the input shaft 10 / the rotational speed of the second output shaft 14).

Here, a case where α1 is 0.43, α2 is 0.44, i1 is 2.12, i2 is 0.92, and i3 is 1.083 is illustrated for calculation of each gear ratio.
In order to simplify the display and calculation formula, α1 · i1 (1 + α2) / {α2 · i3 + α1 · i1 (1 + α2)} is defined as A. In the above-mentioned tooth number ratio, A is 0.734.

First, the speed ratio of the forward first speed (1st) is determined by connecting the input shaft 10 and the first ring gear 30 by engaging the first clutch 50 (C-1) and the second clutch 56 (C-2), and second It is obtained by fixing the second carrier 48 to the case 64 by fastening the brake 70 (B-2).
The speed ratio of the first speed is i1 {α2 · i3 (1 + α1) + α1 · i2} / (α2 · i2 · i3), and the gear ratio set to the above value is 5.208.

Next, the shift to the second speed (2nd) is performed by releasing the engagement of the second brake 70 and maintaining the engagement of the first clutch 50 and the second clutch 56 at the first speed and the first brake 62 (B The first sun gear 30 is fixed to the case 64 by the fastening of -1).
The gear ratio is i1 (1 + α1) / i2, and is 3.295 in the above-mentioned gear ratio.

Next, the shift to the third speed (3rd) is performed by releasing the first brake 62 and maintaining the third clutch 58 (C-3 while maintaining the engagement of the first clutch 50 and the second clutch 56 at the second speed. ).
As a result, the rotating members of the first planetary gear set 24 rotate as a unit, and the gear ratio becomes i1 / i2. In the above gear ratio, the gear ratio is 2.304.

Next, in shifting to the fourth speed (4th), the third clutch 58 is disengaged while the first clutch 50 and the second clutch 56 are kept engaged in the third speed, and the fifth clutch 68 (C -5).
As a result, the gear ratio is i1 · i3 (1 + α1) / [i2 {i3 (1 + α1) + α1 (i1−i3)}]. In the above gear ratio, the gear ratio is 1.789.

Next, the shift to the fifth speed (5th) is performed by releasing the engagement of the fifth clutch 68 and maintaining the engagement of the first clutch 50 and the second clutch 56 at the fourth speed and the fourth clutch 66 (C -4).
As a result, the gear ratio becomes A / (1 + α2) + i1 (1−A) (1 + α1) / i2. In the above gear ratio, the gear ratio is 1.387.

Next, the shift to the sixth speed (6th) is performed by releasing the engagement of the first clutch 50 and the fourth clutch 66 while maintaining the engagement of the second clutch 56 at the fifth speed, and again the third clutch. 58 and the fifth clutch 68 are engaged.
As a result, the connection between the input shaft 10 and the first planetary gear set 24 is released, the first planetary gear set 24 is integrated, and the second sun gear 40 is input via the second gear set (gears 60a, 60b). It is connected to the shaft 10 and the gear ratio becomes i3 / i2. In the above gear ratio, the gear ratio is 1.177.

Next, in the shift to the seventh speed (7th), the engagement of the fifth clutch 68 is released and the fourth clutch 66 is again engaged while maintaining the engagement of the second clutch 56 and the third clutch 58 at the sixth speed. It is done by fastening.
As a result, the gear ratio becomes 1 / (1 + α2) + α2 · i3 / {i2 (1 + α2)}. In the above gear ratio, the gear ratio is 1.054.

Next, in the shift to the eighth speed (8th), the second clutch 56 and the third clutch 58 are disengaged while the fourth clutch 66 is maintained in the seventh speed, and the fifth clutch 68 is again engaged. It is done by fastening. At this time, the first clutch 50 is also engaged but is not involved in power transmission as described above.
As a result, the second planetary gear set 26 is integrated and the second output shaft 14 is directly connected to the input shaft 10, and the gear ratio is 1 regardless of the gear ratio.

Next, in shifting to the ninth speed (9th), the fifth clutch 68 is released and the third clutch 58 is released again while maintaining the engagement of the first clutch 50 and the fourth clutch 66 at the eighth speed. It is done by fastening.
Thereby, the gear ratio becomes i1 / {i1 + α2 (i1-i3)}. In the above gear ratio, the gear ratio is an increase of 0.823.

Next, in shifting to the 10th speed (10th), the engagement of the third clutch 58 is released and the first brake 62 is applied again while maintaining the engagement of the first clutch 50 and the fourth clutch 66 at the ninth speed. It is done by fastening.
As a result, the gear ratio becomes 1 / (1 + α2). In the above gear ratio, the gear ratio is an increase of 0.694.

Next, in the reverse first speed (R-1) in the R range, the first clutch 50 and the third clutch 58 are engaged after the second brake 70 is engaged and the second carrier 48 is fixed. Done in
As a result, the gear ratio becomes −i1 / (α2 · i3). In the above gear ratio, the gear ratio is a reverse rotation of -4.449.

Next, the reverse second speed (R-2) is changed by releasing the third clutch 58 and engaging the fifth clutch 68 while the second brake 70 is engaged.
As a result, the gear ratio becomes -1 / α2. In the above gear ratio, the gear ratio is -2.273.

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 5.208 (1.581)
Second speed 3.295 (1.430)
3rd speed 2.304 (1.288)
4th speed 1.789 (1.290)
5th speed 1.387 (1.178)
6th speed 1.177 (1.117)
7th speed 1.054 (1.054)
8th speed 1.000 (1.215)
9th Speed 0.823 (1.186)
10th speed 0.694

From this, it can be seen that the interstage ratio between the sixth speed and the seventh speed and between the seventh speed and the eighth speed is small. Therefore, if the seventh speed is skipped and the speed is changed from the sixth speed to the eighth speed, the step ratio between the sixth speed and the eighth speed becomes 1.177. In this way, a gear ratio of 9 steps, which is a preferable gear ratio for an automobile, can be obtained.
In this case, although the vehicle normally travels using a transmission gear ratio of 9 forward speeds, as an application example, it is accelerated when shifting from the 1st speed to the 6th speed and then moving to a constant speed. Since it is conceivable to change the speed by selecting either the seventh speed or the eighth speed according to the above, the seventh speed is never wasted.

Thus, it is possible to travel by selecting a more suitable gear ratio by utilizing the gear ratio of 10 forward speeds.
These are the first planetary gear set 24 in a state where the input shaft 10 and the first planetary gear set 24 are not connected by increasing the number of clutches by one compared to the conventional forward eight-stage planetary gear train. Thus, a planetary gear train having 10 forward speeds can be obtained by obtaining a gear ratio of the seventh speed and the eighth speed.

  In the above description, the gear ratio i3 of the second gear set (the gears 60a and 60b) has been described as a reduction ratio of 1.083, but this may be a speed increase ratio smaller than 1. In this case as well, the calculation formulas for the respective gear ratios are the same as described above, but since the gear ratio value changes, the order of the gear positions changes, and the sixth speed and the eighth speed are switched.

In addition, the number of clutches is increased by one compared to the conventional eight-stage planetary gear train. However, as shown in FIG. 2, there are always four friction elements (clutches and brakes) that are not engaged at each gear. The number is the same as the eight-stage planetary gear train in the conventional example.
The frictional elements that are not fastened at each gear stage are factors that reduce the power transmission efficiency of the planetary gear train due to the drag resistance. However, the same number means that the reduction in power transmission efficiency due to the tenth gear is minimized. It means that it can be suppressed.
Therefore, it is possible to improve the acceleration performance and fuel consumption of the vehicle by obtaining the 10 speed ratio and always selecting the optimal speed ratio for traveling.

FIG. 3 shows a skeleton diagram of a multi-stage planetary gear train according to the second embodiment of the present invention, and corresponds to FIG. FIG. 4 shows an operation table corresponding to FIG.
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 first difference between the second embodiment and the first embodiment is that the second carrier 48 can be connected to the sleeve 72 via the fourth clutch 66, and the sleeve 72 is moved to the left and right in the axial direction by an actuator (not shown). Is to be able to.
That is, the sleeve 72 can be connected to the gear 52a of the reduction gear pair integral with the input shaft 10 when moving to the left in the figure, and can be connected to the case 64 when moving to the right.

Accordingly, when the fourth clutch 66 is engaged after moving the sleeve 72 to the left, the input shaft 10 and the second carrier 48 are connected, and when the fourth clutch 66 is engaged after moving the sleeve 72 to the right, the second clutch 66 is engaged. The carrier 48 is fixed to the case 64.
Therefore, the fourth clutch 66 also serves as the function of the second brake 70 in the first embodiment, and the number of friction elements is reduced by one as compared with the first embodiment.

  The second difference is that the first clutch 50 is disposed coaxially with the first output shaft 12 and is provided between the gear 52b of the reduction gear pair and the intermediate shaft 54.

Next, the operation of the second embodiment will be described with reference to the operation table shown in FIG.
The operation table of FIG. 4 is the same as that of the first embodiment shown in FIG. 2 except for the column in which the sleeve 72 is displayed as “S”, and the column of S represents the moving direction of the sleeve 72 with arrows indicating left and right. Yes.
In other words, the sleeve 72 moves to the right in the first forward speed and the reverse speed, and in combination with the engagement of the fourth clutch 66, the second carrier 48 is fixed to the case 64, so that the second carrier 48 is in a higher gear stage than the fifth forward speed. As a result, the second carrier 48 can be connected to the input shaft 10 in combination with the engagement of the fourth clutch 66 by moving to the left side.
An arrow enclosed in parentheses indicates that the sleeve 72 is moving but is not involved in power transmission.
In addition, “*” mark is written in the “S” column of the third speed. This indicates that the left / right movement of the sleeve is switched in accordance with the next expected shift at the third speed, but the movement operation is not limited to the third speed but performed at the second speed or the fourth speed. Also good.

  In the second embodiment, the forward gear ratio and the reverse gear ratio can be obtained by switching the engagement of the friction elements according to the operation table shown in FIG. Since the calculation formulas for the respective gear ratios are the same as those described in the first embodiment, description thereof is omitted.

As described above, also in the second embodiment, it is possible to travel by selecting a more suitable gear ratio by utilizing the gear ratio of 10 forward speeds.
Since the first planetary gear set 24 can be integrated in a state where the input shaft 10 and the first planetary gear set 24 are not connected, the gear ratios of the seventh speed and the eighth speed are obtained. Thus, a planetary gear train having 10 forward speeds can be realized.

  Further, as described above, the number of friction elements is one less than that in the first embodiment. This means that the number of friction elements that are not fastened at each shift stage is small. In general, the frictional elements that are not engaged are factors that cause the drag resistance to deteriorate the power transmission efficiency. In particular, since the second brake 70 in the first embodiment has a large torque capacity, its influence is large. Since the planetary gear train of the second embodiment substantially eliminates the second brake 70 in the first embodiment, the power transmission efficiency is improved.

FIG. 5 shows a skeleton diagram of a multi-stage planetary gear train according to the third embodiment of the present invention, and corresponds to FIG. FIG. 6 shows the positional relationship of each axis as viewed from the left side of FIG.
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 first difference between the third embodiment and the first embodiment is that the second output shaft 14 is arranged in parallel with the input shaft 10 together with the second planetary gear set 26.
The first connecting gear 12 a integrated with the first output shaft 12 and the second connecting gear 14 a integrated with the second output shaft 14 mesh with the output gear 16.
In FIG. 5, the second connecting gear 14a and the output gear 16 are drawn apart from each other, but in actuality they are meshed as shown in FIG. Therefore, the first output shaft 12 and the second output shaft 14 are connected via the first connecting gear 12a, the second linkage gear 14a, and the output gear 16.
In FIG. 6, the first connecting gear 12a, the second linkage gear 14a, and the output gear 16 are drawn with solid lines.

The second difference from the first embodiment is that the first sun gear 30 and the second sun gear 40 are connected by the second gear sets 60a, 60b, and 60c. That is, the intermediate second gear 60 c is arranged coaxially with the input shaft 10, and the input shaft and the tenth second gear 60 c can be connected by the fifth clutch 68.
For this reason, the input shaft 10 is connected to the first sun gear and the second sun gear through the fifth clutch 68 and the second gear sets 60c and 60a, 60b.

The third difference from the first embodiment is that the input shaft 10 can be connected to the second carrier 48 via the transmission gear pair (gears 74 a and 74 b) and the fourth clutch 66.
In addition, although detailed description is omitted, the connection relationship between the rotating members is basically the same as that of the first embodiment.
In FIG. 6, the reduction gear pairs 52a and 52b are drawn with broken lines, and the second gear set (gears 60a, 60b, 60c) is shown with a thick dotted line.

Subsequently, the operation of the fourth embodiment will be described.
Since the operation table of the fourth embodiment is the same as that of the first embodiment shown in FIG. 2, the detailed description is omitted, but the calculation formula of the gear ratio is as follows. Each tooth ratio is defined as follows.
The gear ratio of the first gear set (the number of teeth of the coupling gear 12a / the number of teeth of the coupling gear 14a) is i3, the gear ratio of the transmission gear pair (the number of teeth of the gear 74b / the number of teeth of the gear 74a) is i4, Of the two gear sets, the gear ratio between the gear 60a and the gear 60c (the number of teeth of the gear 60a / the number of teeth of the gear 60c) is i5, and the gear ratio between the gear 60b and the gear 60c (the number of teeth of the gear 60b / the gear 60c). The number of teeth) is i6.
The transmission ratio is represented by the ratio of the rotational speed of the input shaft 10 to the rotational speed of the first output shaft 12 (the rotational speed of the input shaft 10 / the rotational speed of the first output shaft 12). For example, the gear ratio viewed from the second output shaft 14 is also the same.
Further, α2 · i4 · i5 / {i1 · i6 · α1 (1 + α2)} is calculated as B.

Forward 1st speed: i1 {1 + α1 + α1 · i3 · i6 / (i5 · α2)}
Second speed: i1 (1 + α1)
3rd speed: i1
4th speed: i1 · i5 (1 + α1) / {i5 (1 + α1) + α1 (i1-i5)}
5th speed: {i4 + B · i1 (1 + α1) (1 + α2)} / {(1 + B) (1 + α2)}
6th speed: i5
7th speed: (i3 · i6 + α2 · i5) / {i6 (1 + α2)}
8th speed: i3 ・ i4
9th speed: i1 · i3 / {i1 + α2 (i1−i5 / i6)}
10th speed: i3 · i4 / (1 + α2)
Reverse 1st speed: i1, i3, i6 / (α2, i5)
Second gear: i3 / i6 / α2

Although illustration of specific gear ratios is omitted, in the third embodiment as well, in the same way as in the first embodiment, the vehicle is accelerated by obtaining a gear ratio of 10 forward speeds and always selecting the optimum gear ratio for traveling. It becomes possible to improve performance and fuel consumption.
This makes it possible to integrate the first planetary gear set 24 in a state where the input shaft 10 and the first planetary gear set 24 are not connected, as compared with the conventional forward eight-stage planetary gear train. Therefore, a planetary gear train having 10 forward speeds can be obtained by obtaining a gear ratio of the seventh speed and the eighth speed.

  As described above, the multi-stage planetary gear train according to each embodiment of the present invention can obtain a gear ratio of 10 forward and 2 reverse gears preferable for an automobile, and has one clutch compared to the conventional 8-stage planetary gear train. Despite the increase, the reduction in power transmission efficiency due to this can be minimized, so it is possible to improve the acceleration performance and fuel consumption of the vehicle by always selecting the optimal gear ratio and traveling .

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.
Further, as is generally done with an automatic transmission, a one-way clutch can be provided in parallel with the second brake 70 to facilitate shift control from the first speed to the second speed.
Furthermore, although the reduction gear pair (gears 52a and 52b) is composed of two gears, another method having a reduction action and a transmission function between two axes may be used.

In the above description, the second planetary gear set 26 is an example of a single pinion type constituted by the second sun gear 40, the second ring gear 42, the plurality of second pinions 44, and the second carrier 48. A similar function can be obtained as a so-called double pion type planetary gear set.
When the second planetary gear set 26 is a double pion type planetary gear set, the second ring gear 42 and the second carrier 48 may be interchanged with respect to the above description.

  Applicable to vehicles with a smaller vehicle width, especially when applied to a transmission mounted on a horizontally mounted engine vehicle, because the transmission ratio exceeding 8 forward stages can be obtained and the axial length of the transmission part can be shortened. This is possible and can be widely applied to passenger cars and the like where fuel efficiency is important.

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. It is the skeleton figure which showed the multi-stage planetary gear train of this invention. (Example 3) It is a figure which shows arrangement | positioning of the axis | shaft in the multi-stage planetary gear train of Example 3. FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine 2 Torque converter 10 Input shaft 12 1st output shaft 14 2nd output shaft 16 Intermediate output shaft 18 Connection gear 20 Differential gear 22 Axle 24 1st planetary gear set 26 2nd planetary gear set 28 Connection of 1st gear set Gear 30 First Sun Gear 32 First Ring Gear 34 First Pinion 38 First Carrier 40 Second Sun Gear 42 Second Ring Gear 44 Second Pinion 48 Second Carrier 50 First Clutch 52 Gears of Reduction Gear Pair 54 Intermediate Shaft 56 Second Clutch 58 Third clutch 60 Second gear set gear 62 First brake 64 Case 66 Fourth clutch 68 Fifth clutch 70 Second brake 72 Sleeve 74 Transmission gear pair gear

Claims (3)

An input shaft;
A first output shaft provided parallel to the input shaft;
A second output shaft provided coaxially with the input shaft or in parallel with the input shaft and the first output shaft;
A first pinion that is arranged coaxially with the first output shaft and that is engaged with the first sun gear, the first ring gear, the first ring gear, and the first sun gear as rotation members, and a first carrier that rotatably supports the first pinion. A first planetary gear set,
A second sun gear, a second ring gear, a second ring gear and a second pinion meshing with the second sun gear, which are arranged coaxially with the second output shaft, and a second carrier which rotatably supports the second pinion. A second planetary gear set
The input shaft can be connected to the first sun gear and the first ring gear via a reduction mechanism, respectively, and can be connected to the second sun gear and the second carrier, either directly or via a transmission gear,
The first output shaft and the second output shaft are connected to each other via a first gear set,
The first output shaft is connected to the first carrier;
The second output shaft is connected to the second ring gear;
The first sun gear and the second sun gear are connected via a second gear set and can be fixed to a stationary part,
The second carrier can be fixed to the stationary part;
A multi-speed planetary gear configured so that the rotating members of the first planetary gear set can rotate together in a state where the input shaft is not connected to the first sun gear and the first ring gear. Column.   The second output shaft is provided coaxially with the input shaft, an intermediate output shaft is provided in parallel with the input shaft and the second output shaft, and from one gear of the connecting gear integrated with the second output shaft, The multi-speed planetary gear train according to claim 1, wherein the intermediate output shaft is driven.   The second output shaft is provided in parallel with the input shaft, and the first drive gear integral with the first output shaft and the second drive gear integral with the second output shaft are both engaged with the output gear. The multi-speed planetary gear train according to claim 1, wherein the multi-speed planetary gear train is configured.

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