JP2009210087A - Multistage shift planetary gear train - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve power transmission efficiency to improve fuel economy, in a multistage shift planetary gear train of forward eight stages or more. <P>SOLUTION: This multistage shift planetary gear train includes an input shaft 10, an output shaft 12, and four sets of single pinion type planetary gear sets 14, 16, 18a and 18b. The input shaft 10 is connected to a second carrier 38 and connectable to a fourth sun gear 50, and the output shaft 12 is connected to a third ring gear 42. A first sun gear 20 is connected to a second sun gear 30 and fixable to a stationary part 64, and a first ring gear 22 can be fixed to the stationary part 64. A first carrier 28 is connected to a third carrier 48 and a fourth carrier 58. A second ring gear 32 can be connected to a third sun gear 40 and a fourth ring gear 52, and can be connected to a fourth sun gear. The third sun gear 40 can be connected to either of the second ring gear 32 and the input shaft 10. <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 a planetary gear train used for an automatic transmission for vehicles, a gear train capable of multi-speed shifting of eight forward speeds has been proposed 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 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 conventional planetary gear train requires a double pinion type planetary gear connected to the input shaft in order to obtain a preferable gear ratio for an automobile in eight forward speeds, it is compared with a so-called single pinion type planetary gear. The manufacturing cost is high, and there is a problem that the power transmission efficiency is poor due to the large meshing of the gears. On the other hand, four of the six friction elements always rotate freely. The drag torque of the vehicle is increased, and coupled with the problem that 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 a double pinion type planetary gear with poor power transmission efficiency is required and four friction elements are always idle, so that the fuel consumption is poor and the heat generation is large.
An object of the present invention is to improve fuel efficiency and heat generation by using a single pinion type planetary gear and reducing the number of friction elements that are always idle while ensuring a preferable gear ratio for a vehicle such as an automobile. The object is to provide a possible multi-speed planetary gear train.

  The multi-stage planetary gear train of the present invention includes an input shaft, an output shaft provided coaxially with the input shaft, and a coaxial arrangement with the input shaft. The first sun gear, the first ring gear, the first ring gear, and the first sun gear A first planetary gear set having a meshed first pinion, a first carrier that rotatably supports the first pinion, a second sun gear, a second ring gear, a second gear engaged with the second ring gear, and a second sun gear; 2 pinion, a second planetary gear set having a second carrier that rotatably supports the second pinion, and coaxially arranged with the output shaft, the input member, the output member, and the torque from the input member to the output member A reaction force torque smaller than the reaction force torque received by the first reaction force member when the torque is transmitted from the input member to the output member. A planetary gear group having a second reaction force member, an input shaft connected to the second carrier and connectable to the input member, an output shaft connected to the output member, and the first sun gear It can be connected to the second sun gear and fixed to the stationary part, the first ring gear can be fixed to the stationary part, the first carrier can be connected to the first reaction force member, and the second ring gear can be connected to the input member The second reaction force member can be connected to either the second ring gear or the input shaft.

  Since the multi-speed planetary gear train of the present invention is configured as described above, a single pinion type with high power transmission efficiency can be used while obtaining a preferable gear ratio for a transmission for an automobile or the like, and it can always rotate freely. Since the number of friction elements can be two, two less than the conventional example, 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 18a, and the fourth planetary gear set 18b, which are sequentially arranged from the upstream side to the downstream side, are generally single pinion types. They are called, 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 configured as rotating members. The fourth planetary gear set 18b includes a fourth sun gear 50, a fourth ring gear 52, and a plurality of second gears. It is composed of rotating members such as a 4-pinion 54 and a fourth carrier 58.

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 fourth sun gear 50 is the input member of the present invention, the third ring gear 42 is the output member of the present invention, the third carrier 48 and the fourth carrier 58 are connected, and the first reaction force member of the present invention is connected. The 3rd sun gear 40 and the 4th ring gear 52 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 input shaft 10 is connected to the second carrier 38 and can be connected to the fourth sun gear 50 that is the input member of the present invention by the first clutch 60.
The first sun gear 20 is connected to the second sun gear 30 and can be fixed to a transmission case (stationary portion) 64 by a first brake 62.

The first ring gear 22 can be fixed to the case 64 by the second brake 66.
The 1st carrier 28 is connected with the 3rd career 48 and the 4th career 58 which are the 1st reaction force members of the present invention.
The second ring gear 32 can be connected to the fourth sun gear 50 that is the input member of the present invention by the second clutch 68, and the third sun gear 40 and the second sun gear 40 that are the second reaction force member of the present invention by the third clutch 70. It can be connected to the 4-ring gear 52.
The third ring gear 42 that is the output member of the present invention is connected 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 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 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, the calculation of each gear ratio is illustrated for the case where the respective gear ratio is 0.48, α2 is 0.54, α3 is 0.31, and α4 is 0.62.
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, (1 + 1 / α4) / α2 will be described as A.
In the above-mentioned tooth number ratio, A is 4.839.

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 66 ( B-2) is performed. In subsequent shifts, the engagement of the second brake 66 is maintained up to the fifth speed.
The speed ratio of the first speed is 1 / α3 · α4, and the gear ratio set to the above value is 5.203.

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 68 (C-2).
The speed ratio of the second speed is 1 / {α3 · α4 (1 + α2)}, which is 3.379 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 + α1) / (α1 + α3 · α4).
In the above-mentioned tooth number ratio, it is 2.202.

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 70 (C-3).
The speed ratio of the fourth speed is {α1 (1 + α2) + α2} / {α1 (1 + α2)}. In the above-mentioned tooth number ratio, it is 1.731.

Next, the shift to the fifth speed (5th) is performed by releasing the engagement of the second clutch 68 at the fourth speed and engaging the first clutch 60.
The speed ratio of the fifth speed is {1 + α1 (1 + α4 · A)} / {α3 · α4 + α1 (1 + α4 · A)}. In the above-mentioned tooth number ratio, it is 1.382.

Next, the shift to the sixth speed (6th) is performed by releasing the engagement of the second brake 66 up to the fifth speed and engaging the first clutch 60 again.
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 68 at the sixth speed and engaging the first brake 62.
As a result, the gear ratio is A / (A + 1 / α4-α3), and the above gear ratio is 0.788.

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 68 again.
As a result, the gear ratio becomes 1 / (1 + α2), and the speed increases by 0.649 in the aforementioned gear ratio.

Next, the reverse shift of the R range is performed by engaging the third clutch 70, the first brake 62, and the second brake 66.
As a result, the transmission gear ratio becomes -1 / [alpha] 3 (1+ [alpha] 2), and in the above-described gear ratio, -2.095.

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 5.203 (1.540)
2nd speed 3.379 (1.534)
3rd speed 2.202 (1.272)
4th speed 1.731 (1.252)
5th speed 1.382 (1.382)
6th speed 1.000 (1.269)
7th speed 0.788 (1.213)
8th 0.649

From this, it can be seen that a gear ratio of 8 steps, which is a preferable gear ratio for an automobile, can be obtained.
In addition, since the planetary gear group 18 has the above-described configuration, the gear ratio range (the gear ratio of the first speed / the gear ratio of the eighth speed) can be as large as 8.012. It is suitable for transmissions that require a large gear ratio range, such as cars.

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. For the gear sets 14, 16, 18a, 18b, a single pinion type having a simple structure, light weight and high power transmission efficiency can be used, and the number of friction elements constantly spinning is two. Since there are two less than the example, drag resistance (drag torque) of the friction element that is 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. 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 connection relationship between the third planetary gear set 18a and the fourth planetary gear set 18b constituting the planetary gear group 18 is different.
That is, the fourth ring gear 52 is the input member of the present invention, the third ring gear 42 and the fourth carrier 58 are connected to each other to output the present invention, the third carrier 48 is the first reaction force member of the present invention, The third sun gear 40 and the fourth sun gear 50 are connected to constitute the second reaction force member of the present invention.
The connection relationship between these members and the rotating members of the first planetary gear set 14 and the second planetary gear set 16 is the same as that of the first embodiment, and the description thereof is omitted.

Next, the operation of the second embodiment will be described.
The operation table of the second embodiment is the same as that of the first embodiment shown in FIG. 2, and the operation of the friction element at each shift is the same as that of the first embodiment, and thus detailed description thereof is omitted.
Although description of the calculation of each gear ratio is omitted, in the second embodiment as well, in the same manner as in the first embodiment, it is possible to perform eight forward and reverse shifts.
Although the possible gear ratio range is smaller than that of the first embodiment, the input member of the present invention is the fourth ring gear 52, and therefore the radius on which the input torque to the input member acts is large, and the planetary gear group 18 at the forward low speed stage. There is a merit that the tooth root stress is reduced.

As in the first embodiment, the multi-stage planetary gear train according to the second embodiment of the present invention can obtain a preferable gear ratio for the automobile in the eight forward speeds and the first reverse speed, and the gear train is configured as described above. The planetary gear sets 14, 16, 18 a, and 18 b can use a single-pinion type that has a simple structure, is light in weight, and has high power transmission efficiency, and has two friction elements that are always idle. Since there are two less than the conventional example, drag resistance of the friction element that is 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.
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 connection relationship between the third planetary gear set 18a and the fourth planetary gear set 18b constituting the planetary gear group 18 is different.
That is, the point that the fourth sun gear 50 constitutes the input member of the present invention is the same as that of the first embodiment, but the third ring gear 42 and the fourth carrier 58 are connected to connect the output member of the present invention to the third carrier. 48 and the fourth ring gear 52 are connected to constitute a first reaction force member of the present invention, and the third sun gear 40 constitutes a second reaction force member of the present invention.
The connection relationship between these members and the rotating members of the first planetary gear set 14 and the second planetary gear set 16 is the same as that of the first embodiment, and the description thereof is omitted.

Next, the operation of the third embodiment will be described.
The operation table of the third embodiment is the same as that of the first embodiment shown in FIG. 2, and the operation of the friction element at each speed change is also the same as that of the first embodiment, so that detailed description is omitted.
Although explanation of the calculation of each gear ratio is omitted, in the third embodiment, as in the first embodiment, it is possible to perform eight forward speeds and reverse gears.
Although the possible gear ratio range is slightly smaller than that of the first embodiment, it can cover the range of gear ratios widely used in passenger cars.

Similarly to the first embodiment, the multi-stage planetary gear train according to the third embodiment of the present invention can obtain a preferable gear ratio for the automobile in the eight forward speeds and the first reverse speed, and the gear train is configured as described above. The planetary gear sets 14, 16, 18 a, and 18 b can use a single-pinion type that has a simple structure, is light in weight, and has high power transmission efficiency, and has two friction elements that are always idle. Since there are two less than the conventional example, drag resistance of the friction element that is 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.
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 fourth embodiment and the first embodiment is that the arrangement of the third planetary gear set 18a and the fourth planetary gear set 18b constituting the planetary gear group 18 is switched first, and the second reaction force is second. The third sun gear 40 and the fourth ring gear 52 constituting the member are connectable to the input shaft 10 by the third clutch 70, and thirdly, the third carrier 48 and the second carrier member constituting the second reaction force member The fourth carrier 58 can be connected to the second ring gear 32 by the fourth clutch 72, and fourthly, a one-way clutch (OWC) 74 is provided in parallel with the first brake 62.

  In other words, the first difference between the third planetary gear set 18a and the fourth planetary gear set 18b, which is the first difference, is merely an arrangement problem, and the first clutch 60 and the second clutch are related to this. 68. Although the arrangement of the third clutch 70 is different from that of the first embodiment, it does not affect the operation described later.

  Since the third sun gear 40 and the fourth ring gear 52 constituting the first reaction force member, which is the second difference, can be connected to the input shaft 10 by the third clutch 70, the speed change particularly with the reverse gear ratio. A slight difference occurs in the ratio calculation formula and operation table, as described below.

  Since the third carrier 48 and the fourth carrier 58 constituting the second reaction force member, which is the third difference, can be connected to the second ring gear 32 by the fourth clutch 72, the number of shift stages is increased as described later.

  The one-way clutch (OWC) 74 provided in parallel with the first brake 62 as the fourth difference is that the first sun gear 20 and the second sun gear 30 are attached to the case 64 only when torque in the direction of accelerating the vehicle is applied. It has a function of fixing, and also has a function of fixing the first carrier 28 to the case 64 together with the second brake 66. This facilitates shift control as will be described later.

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.
The operation table of FIG. 6 is written in the same manner as in the first embodiment shown in FIG. 2, but in addition to the increase in the number of shift stages, a one-way clutch (OWC) 74 is provided. The operation is also added.

In addition, in the calculation of each gear ratio in the following description, the gear ratio of each planetary gear set 14, 16, 18a, 18b is respectively α1, 0.62, α2, 0.56, α3, 0.43, An example in which α4 is set to 0.49 will be described.
In order to simplify the calculation formula, α1 {1 / α3 + 1 / (α2 · α3)} will be described as B.
In the above-mentioned tooth number ratio, B is 3.525.

As described in the first embodiment, the speed change is performed by operating the friction elements and the one-way clutch 74 in accordance with the operation table shown in FIG.
Although a detailed description of the operation of each friction element is omitted, the one-way clutch 74 works together with the second brake 66 to accelerate the vehicle at the first speed and the second speed, and the first sun gear 20 and the second sun gear. 30 is fixed to the case 64 and is automatically released when shifting to the third speed.

For this reason, at the first speed and the second speed in the D range, the input shaft 10 cannot be driven from the output shaft 12 side as in the so-called engine brake.
As indicated by the L range, at the first speed and the second speed when the first brake 62 is engaged, the input shaft 10 can be driven from the output shaft 12 side, and the engine brake can be applied.

  Further, since the entire planetary gear train is integrated at the eighth speed, the transmission ratio is 1 regardless of the gear ratio, but the engagement of the friction element at the eighth speed is the first to fourth as shown in the operation table. Three of the clutches 60, 68, 70, 72 may be fastened, or all of them may be fastened, and the fastening of only two of the first clutch 60 and the second clutch 68 can be integrated. .

The formula for calculating the gear ratio of each gear stage is as follows.
1st speed: 1 / (α3 ・ α4)
Second speed: 1 / {α3 · α4 (1 + α2)}
3rd speed: {α2 + α1 (1 + α2)} / {α1 (1 + α2) −α2 · α3}
Fourth gear: (1 + α1) / (α1 + α3 · α4)
5th speed: {α1 (1 + α2) + α2} / {α1 (1 + α2)}
6th speed: (B + 1 / α4) / (B + α3)
7th speed: {α2 · α4 + α1 (1 + α2 · α4 + α4)} / {α1 (1 + α2 · α4 + α4)-α2 · α3 · α4}
8th speed: 1
9th speed: (1 + α4) / {1 + α4 (1 + α2) + α2 ・ α3 ・ α4}
10th speed: 1 / (1 + α2-α2, α3, α4)
11th speed: 1 / (1 + α2)
12th speed: 1 / (1 + α2 + α2 ・ α3)
Reverse: -1 / α3

Based on these calculation formulas, the forward gear ratio in the case of the above-mentioned gear ratio is arranged as follows. The numerical value in the parenthesis on the right side is the ratio (interstage ratio) with the gear ratio higher by one stage.
1st speed 4.746 (1.560)
2nd speed 3.042 (1.447)
3rd speed 2.102 (1.078)
4th speed 1.950 (1.235)
5th speed 1.579 (1.122)
6th speed 1.407 (1.004)
7th speed 1.402 (1.402)
8th speed 1.000 (1.263)
9th speed 0.792 (1.141)
10th speed 0.693 (1.081)
11th speed 0.641 (1.155)
12th speed 0.555

Looking at these gear ratios, there are gear speeds with a small gear ratio. For example, the sixth gear and the seventh gear are reversed in the magnitude ratio of the gear ratio depending on the gear ratio, but usually one of them may be used.
In addition, it is conceivable that another gear stage having a small inter-step ratio performs a one-step or two-step jumping shift.
As an example, in normal acceleration, the gear ratio and the step ratio are arranged as follows assuming that the gears are shifted over the third speed, the fifth speed, and the seventh speed.
1st speed 4.746 (1.560)
2nd speed 3.042 (1.560)
Fourth speed 1.950 (1.391)
6th speed 1.407 (1.407)
8th speed 1.000 (1.263)
9th speed 0.792 (1.141)
10th speed 0.693 (1.081)
11th speed 0.641 (1.155)
12th speed 0.555

If it does in this way, it will become a transmission gear ratio of 9 steps suitable for a transmission for vehicles. However, when shifting from the acceleration state to the steady running, it is possible to select the optimum gear ratio for the running condition from among the 12 gear ratios, and therefore the skipped gear ratio is not necessarily used. Depending on the conditions of use, it may be used, for example, as a 10-speed transmission, and in this case as well, each speed ratio and the inter-stage ratio can be used well.
Thus, this embodiment is suitable for use as a transmission having 9 to 10 forward gears.

  In addition, since the third sun gear 40 and the fourth ring gear 52 constituting the first reaction force member can be connected to the input shaft 10 by the third clutch 70, particularly the gear ratio of the planetary gear group 18 is reduced. The reverse gear ratio can be made large without any change, and the gear ratio width (first speed gear ratio / maximum speed gear ratio) can easily be made large so as to exceed 8.5.

Also in the multi-stage planetary gear train according to the fourth embodiment of the present invention, a preferable gear ratio is obtained for the automobile in 9 or more forward speeds and 1 reverse speed, and the gear train is configured as described above. , 16, 18a, and 18b can use a single pinion type with a simple structure, light weight and high power transmission efficiency, and the number of friction elements constantly spinning is three, Since there is one less than the conventional example, drag resistance of the friction element that is idle can be reduced.
As a result, it is possible to obtain a multi-stage automatic transmission with 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 gear ratio that is preferable for automobiles and a forward gear ratio of 9 stages or more, that is excellent in fuel efficiency, and that generates little heat. It becomes possible to obtain.
The planetary gear group 18 can be applied to any planetary gear train other than the combinations described in the first to fourth embodiments as long as it is generally used as a forward three-stage planetary gear train. is there.

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, the arrangement of the first to fourth planetary gear sets 14, 16, 18a, 18b and the friction elements 60, 62, 66, 68, 70, 72 can be changed according to the layout of the transmission. it can.

  Since it is possible to obtain an automatic transmission that has a forward gear ratio of eight or more forward speeds, excellent fuel efficiency, and low heat generation, it can be widely applied to small passenger cars that emphasize fuel efficiency and medium-sized 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 18a 3rd planetary gear set 18b 4th planetary gear set 20 1st sun gear 22 1st ring gear 24 1st Pinion 28 1st carrier 30 2nd sun gear 32 2nd ring gear 34 2nd pinion 38 2nd carrier 40 3rd sun gear 42 3rd ring gear 44 3rd pinion 48 3rd carrier 50 4th sun gear 52 4th ring gear 54 4th pinion 58 Fourth carrier 60 First clutch 62 First brake 64 Case 66 Second brake 68 Second clutch 70 Third clutch 72 Fourth clutch 74 One-way clutch

Claims (6)

An input shaft;
An output shaft;
The first sun gear, the first ring gear, the first ring gear, a first pinion meshing with the first sun gear, and a first carrier that rotatably supports the first pinion are arranged coaxially with the input shaft. A first planetary gear set;
A second planetary gear set including 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 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 second carrier and is connectable to the input member;
The output shaft is connected to the output member;
The first sun gear is connected to the second sun gear and can be fixed to a stationary portion;
The first ring gear can be fixed to a stationary part,
The first carrier is connected to the first reaction force member;
The second ring gear is connectable with the input member;
The multi-stage planetary gear train, wherein the second reaction force member is connectable with either the second ring gear or the input shaft. The planetary gear group 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 planetary gear having a third carrier that rotatably supports the third pinion. A gear set,
A fourth sun gear, a fourth ring gear, a fourth pinion gear meshing with the fourth ring gear and the fourth sun gear, and a fourth planetary gear set including a fourth carrier rotatably supporting the fourth pinion,
The fourth sun gear constitutes the input member, the third ring gear constitutes the output member, the third carrier and the fourth carrier are connected to each other to constitute the first reaction force member, The multi-stage planetary gear train according to claim 1, wherein a third sun gear and the fourth ring 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 pinion meshed with the third ring gear and the third sun gear, and a third planetary gear provided with a third carrier rotatably supporting the third pinion. A gear set,
A fourth sun gear, a fourth ring gear, a fourth pinion gear meshing with the fourth ring gear and the fourth sun gear, and a fourth planetary gear set including a fourth carrier rotatably supporting the fourth pinion,
The fourth ring gear constitutes the input member, the third ring gear and the fourth carrier are connected to each other to constitute the output member, the third 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 pinion meshed with the third ring gear and the third sun gear, and a third planetary gear having a third carrier that rotatably supports the third pinion. A gear set,
A fourth sun gear, a fourth ring gear, a fourth pinion gear meshing with the fourth ring gear and the fourth sun gear, and a fourth planetary gear set including a fourth carrier rotatably supporting the fourth pinion,
The fourth sun gear forms the input member, the third ring gear and the fourth carrier are connected to each other to form the output member, and the fourth ring gear and the third carrier are connected to each other to 2. The multi-stage 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.   The multi-stage planetary gear train according to any one of claims 1 to 4, further comprising a clutch that connects the second ring gear and the first reaction force member.   6. The multi-speed planetary gear train according to claim 1, wherein a first brake for fixing the first sun gear to a stationary portion is provided, and a one-way clutch is provided in parallel therewith.

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