JP2010002028A - Friction element of multi-stage shift planetary gear train - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To heighten the power transmission efficiency to improve fuel cost in a 8-forward multi-stage shift planetary gear train. <P>SOLUTION: This multi-stage shift planetary gear train is disposed between an input shaft 10 and an output shaft 12. Friction elements 64, 66, 68, 114, 116 are capable of fixing a first reaction force member 52 of the gear train to a static part 63, or connecting an output member 112 and an output shaft 12 to each other. The first friction elements 64, 114 are connected in the forward first to fifth speed movement and backward movement. The second friction elements 66, 116 are disposed parallel to the first friction elements. At least in the forward first and second speed movement and backward movement, both the first friction elements 64, 114 and the second friction elements 66, 116 are connected. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

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

A number of planetary gear trains for use in automatic transmissions for vehicles have been proposed, including those of the present applicant, that are capable of multi-speed shifting of eight forward speeds, mainly for the purpose of improving vehicle fuel efficiency, exhaust characteristics, acceleration performance, and the like. ing.
As a conventional planetary gear train capable of such multi-speed shifting, there is a multi-speed planetary gear train composed of four sets of planetary gears and five friction elements, and this gear train is always one of the five friction elements. The forward eight speed ratio is obtained by switching the three friction elements to be engaged. (See Patent Document 1).

However, in the above conventional planetary gear train, the friction element that is fastened in the forward first to fifth speeds and the reverse travel requires a large capacity because a large torque acts particularly at the time of starting, and this friction element is released. In traveling at a high speed stage of forward sixth speed or higher, there is a problem that a large drag resistance (drag torque) is generated because of a large capacity, which causes a deterioration in fuel consumption and heat generation.
Japanese Patent Application No. 2008-150203

The problem to be solved is that the friction elements that are fastened in the forward first to fifth speeds and the reverse speed generate a large drag resistance in traveling at high speeds, resulting in deterioration of fuel consumption and heat generation.
An object of the present invention is to provide a multi-stage planetary gear train capable of achieving eight forward speeds, while the friction elements that are fastened in the first to fifth forward speeds and the reverse speed satisfy a large fastening torque that acts at the time of starting, An object of the present invention is to provide a friction element of a multi-speed planetary gear train that can reduce drag resistance in high-speed traveling of 6 or more forward speeds and reduce fuel consumption and heat generation.

The friction element of the multistage planetary gear train of the present invention is:
The first reaction force member of the multi-speed planetary gear train arranged between the input shaft and the output shaft can be fixed to the stationary portion, or the output member and the output shaft can be connected, and the first forward speed to the fifth forward speed A first friction element fastened and fastened in reverse,
A second friction element arranged in parallel with the first friction element,
Both the first friction element and the second friction element are fastened at least in the forward first speed, the second speed, and the reverse speed.

Since the friction element of the multi-speed planetary gear train of the present invention is configured as described above, in the multi-speed planetary gear train capable of achieving eight forward speeds, the friction element is fastened in the forward first speed to the fifth speed and the reverse speed. However, the large torque acting at the start of the vehicle is satisfied, and the drag resistance in the high-speed traveling of the sixth forward speed or more is reduced, so that the deterioration of fuel consumption and heat generation can be reduced.

  Hereinafter, friction elements of 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 multi-speed planetary gear train and its friction elements according to Embodiment 1 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 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 62.

When the second clutch 62 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 case (stationary portion) 63 of the transmission by the first brake 64 and can also be fixed to the case 63 by the second brake 66.

As will be described later, the second brake 66 is an auxiliary brake having a conical friction surface, and the fourth ring gear 52 is fixed to the transmission case 63 by fastening only the first brake 64, and the first brake 64. And the second brake 66 may be fastened and fixed.
The first sun gear 20 is integrally connected to the third ring gear 42 and can be fixed to the case 63 by the third brake 68.

The output shaft 12 is connected to the fourth carrier 58 and can be connected to the second carrier 38 via the third clutch 70.
Here, the fourth ring gear 52 is the first reaction member of the present invention, the first brake 64 is the first friction element of the present invention, and the second brake 66 is the second friction element of the present invention. Constitute.
The first sun gear 20 and the third ring gear 42 are the second reaction force member of the present invention.

Next, the operation of the planetary gear train and its friction element in 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 64, 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.

Although the detailed description is omitted, the embodiment of the present invention shown in FIG. 1 has eight forward stages by fastening each friction element 60, 62, 64, 66, 68, 70 as shown in FIG. A reverse gear ratio of one reverse can be obtained.
For example, 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 set 14 is α1, the second planetary gear set 16 is α2, the third planetary gear set 18 is α3, the fourth planetary gear set 19 is α4, and the respective gear ratios are α1 of 0.60 and α2. Examples of gear ratios when 0.60, α3 is 0.60, and α4 is 0.55 are as follows.
1st speed 7.515
Second speed 4.697
3rd speed 2.818
4th speed 2.091
5th speed 1.481
6th speed 1.000
7th speed 0.816
8th speed 0.625
Reverse -6.162

Here, torque acting on the first brake 64, the second brake 66, and the third brake 68 will be described. The following numerical values are values when the input torque is 1.
In the case of the forward first speed, the output torque is 7.515 from the above gear ratio, and the difference of -6.515 from the input torque is applied to the first brake 64, the second brake 66, and the third brake 68 as the reaction torque. Works.

At this time, since -1 / α3 torque acts on the third ring gear 42, the torque acting on the third brake 68 is −1.667, which is a difference from the aforementioned −6.515, −4.848. Acts on the fourth ring gear 52.
Similarly, in the case of the second speed, the output torque is 4.697, and the difference -3.697 from the input torque acts on the first brake 64, the second brake 66, and the third brake 68 as the reaction torque.

Among them, the torque acting on the first sun gear 20 and the torque acting on the third ring gear 42 are canceled out and the torque of α1 / (1 + α1) −1 / α2 (1 + α1) acts on the third brake 68. .667, and the difference from -3.697 described above, -3.030 torque, acts on the fourth ring gear 52.
From the third speed onward, the reaction force torque acts only on the fourth ring gear 52, so the reaction force torque acting on the fourth ring gear 52 at each gear stage up to the fifth speed is as follows.
1st speed 4.848
2nd speed 3.030
3rd speed 1.818
4th speed 1.091
5th speed 0.481

From this, it can be seen that the reaction torque acting on the fourth ring gear 52 has a large difference between the first speed and the fifth speed.
In general, in an automatic transmission for an automobile, a multi-plate brake acting by hydraulic pressure is used to fix a member that receives a reaction torque, such as the third ring gear 42 and the fourth ring gear 52, to the case 63. When 52 is fixed to the case 63 with one multi-plate brake, there are the following two problems.

That is, in order to satisfy the reaction force torque of the first speed, a large brake capacity exceeding 4.848 is required, and this rotates freely at the high speed stage of the sixth speed or higher, so that a large drag torque is generated at the high speed stage. Will occur.
The first problem is a deterioration in fuel consumption and heat generation due to a large drag torque in traveling at a high speed stage.
The second problem is that there is a very large brake capacity with respect to the torque required for the fifth and fourth speeds when shifting from the sixth gear or higher to the fifth or fourth speed. Therefore, it is difficult to control so-called shift shock.

There is a merit that the first brake 64 and the second brake 66 are used together to fix the fourth ring gear 52 to the case 63.
That is, both the first brake 64 of the multi-plate brake and the second brake 66 of the conical friction brake are engaged at the first speed and the second speed that require a particularly large brake torque. For this reason, the total capacity of both the first brake 64 and the second brake 66 only needs to satisfy the reaction torque of the first speed.

In general, a conical friction brake can obtain a large friction torque with the same pressing force, but has a characteristic that a drag resistance during idling is small when the capacity is the same.
Assuming that the effective radius of the friction surface is the same, a single conical clutch can obtain about six times the friction torque with respect to one flat multi-plate clutch, and the drag resistance during idle rotation is almost the same. Because it can be said.

  The fact that the friction torque is large for the same pressing force is difficult to control the hydraulic shock by controlling the hydraulic pressure, but as shown in the operation table of FIG. 2, the second brake 66 is engaged / released. Switching is performed when shifting from the 3rd speed or the 4th speed to the 2nd speed, and since the 4th ring gear 52 is engaged in a state where it is fixed, the engagement of the second brake 66 affects the transmission shock. There is no fear of affecting.

Therefore, the first brake 64 needs only to have a capacity that satisfies the reaction torque at the third speed or higher gear, so the required brake torque is 1.818.
When the second brake 66 is not provided, the required brake torque of the first brake 64 is 4.848, so that a capacity of about 40% is sufficient.

Therefore, even if the drag resistance generated by the second brake 66 itself during traveling at a high speed stage is taken into account, the drag resistance related to the brake that fixes the fourth ring gear 52 is reduced to about half by providing the second brake 66. Is possible.
Further, as described above, since the capacity of the first brake 64 is reduced, there is a merit that it becomes easier to control the shift shock when shifting from the sixth speed or the seventh speed to the fifth speed or the fourth speed.

Next, specific structures of the first brake 64 and the second brake 66 will be described with reference to FIG.
3 is a cross-sectional view showing the main parts of the first brake 64, the second brake 66, and the third brake 68 in FIG.
A spline 63a and a first cylinder 63b are formed in the case 63, a first piston 72 is inserted into the first cylinder 63b, and the first piston 72 is illustrated by hydraulic pressure supplied to the first cylinder 63b from an oil passage (not shown). By moving to the middle left side, it is possible to press the multi-plate friction plates 74, 76 provided between the fourth ring gear 52, which shows only a part, and the spline 63 a of the case 63.

  Although detailed description is omitted, the friction plate 74 is integrated with the spline 63a of the case 63 in the rotation direction, and the friction plate 76 is integrated with the spline 52a formed in the fourth ring gear 52 in the rotation direction. A friction material is attached to either one of the friction plates 74 and 76, and the fourth ring gear 52 is fixed to the case 63 by the friction between the friction plates 74 and 76 by being pressed by the first piston 72. Perform the action.

On the other hand, on the left side of the friction plates 74 and 76, there is a disc spring 80 whose end is fixed to the case 63 in the axial direction by a snap ring 78, and the connecting member 82 is provided between the disc spring 80 and the friction plates 74 and 76. And are provided.
Here, when both the friction plates 74 and 76 are pressed by the first piston 72, the disc spring 80 is bent so as to be a flat plate, and the connecting member 82 is left together with the friction plates 74 and 76 by the amount of deformation of the disc spring 80. Will be moved to.

An inner cone 84 is fixed to the left end of the connecting member 82 by two snap rings 86. A conical friction surface 84 a is formed on the outer side of the inner cone 84.
On the left side of FIG. 3, a second cylinder 90 is fixed to the case 63, and a second piston 92 is inserted therein.
Between the second piston 92 and the case 63, multi-plate friction plates 94 and 96 are provided almost symmetrically with the first brake 64 described above, and the third ring gear 42, in which only a part is drawn, is provided with the case 63. It is possible to fix to.

That is, both the friction plates 94 and 96 can be pressed by the second piston 92 moving to the right side in the figure by the hydraulic pressure supplied to the second cylinder 90 from an oil passage (not shown).
Similarly to the description of the first brake 64, a snap ring 98, a connecting member 102, and a disc spring 104 are provided. A shoulder portion 102a is formed on the connecting member 102, and an outer cone 106 that contacts the shoulder portion 102a is a snap ring. 108 is integrated with the connecting member 102 in the axial direction.

The outer cone 106 is formed with a spline 106a meshed with the spline 63a and a conical friction surface 106b. A friction material 106c is affixed to the conical friction surface 106b and corresponds to the conical friction surface 82a of the inner cone 84.
The inner cone 84 and the outer cone 106 constitute a second brake 66.
That is, when a pressing force acts between the conical friction surface 82a and the conical friction surface 106b, the fourth ring gear 52 can be fixed to the case 63 by the friction between the two.

  As described above, the first piston 72 moves to the left side and the second piston 92 moves to the right side, so that the conical friction surface 82a and the conical friction surface 106b are pressed against each other. The dimensions and positions of the related parts are such that the conical friction surface 86a and the conical friction surface 106b do not come into contact with each other only by moving either one of the two pistons 92.

Next, the operation of the first brake 64, the second brake 66, and the third brake 68 shown in FIG. 3 will be described.
As described above, the first brake 64 and the third brake 68 supply hydraulic pressure to the first cylinder 63 b and the second cylinder 90, respectively, so that the first brake 64 independently causes the fourth ring gear 52 and the third brake 68. The third ring gear 42 can be fixed to the case 63.

Therefore, as shown in the operation table of FIG. 2, when one of the fourth ring gear 52 and the third ring gear 42 is fixed to the case 63, the fixing action is performed independently.
On the other hand, when both the first brake 64 and the third brake 68 are fastened together as in the forward first speed, the second speed, and the reverse speed, as described above, between the conical friction surface 86a and the conical friction surface 106b. The fourth ring gear 52 is fixed to the case 63 by the action of both the first brake 64 and the second brake 66.

  Therefore, when large brake torque is required such as forward first speed, second speed and reverse, the total capacity of both the first brake 64 and the second brake 66 only needs to satisfy the required brake torque. As a result, the first brake 64 has a capacity that satisfies the reaction torque of the third speed.

  According to the friction element of the multi-stage planetary gear train of the first embodiment described as described above, the drag torque at the high speed stage of the sixth speed or higher can be reduced correspondingly, and the deterioration of fuel consumption and the temperature increase of the transmission can be reduced. There is an advantage that a shift shock is less likely to occur in the shift from the high speed to the fifth speed.

Further, as described above, the second brake 66 operates in conjunction with the engagement of both the first brake 64 and the third brake 68, and therefore does not require a piston or hydraulic circuit for its own engagement. Thus, the manufacturing cost required to provide the second brake 66 can be suppressed.

FIG. 4 shows a skeleton diagram of the multi-stage planetary gear train and its friction elements 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 planetary gear train is different and the configuration of the friction elements is different.
First, for the planetary gear train, the input shaft 10 and the output shaft 12 are disposed in parallel, and the first planetary gear set 14 disposed on the input shaft 10 includes a first sun gear 20, a first ring gear 22, A plurality of first pinions 24 meshed with the first ring gear 22 and the first sun gear 20, and a rotating member such as a first carrier 28 that rotatably supports the first pinion 24.

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 third planetary gear set 18 is a so-called double pinion type, and is engaged with the third sun gear 40, the third ring gear 42, the outer pinion 44 engaged with the third ring gear 42, the outer pinion 44 and the third sun gear 40. The inner pinion 46 and the third carrier 48 that rotatably supports the outer pinion 44 and the inner pinion 46 are configured.

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 input shaft 10 and the output shaft 12 will be described.
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 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 62.
The second carrier 38 can be connected to the output shaft 12 via the connecting gear pair 110 and the third clutch 70.

The third carrier 48 is connected to the first sun gear 20 and can be fixed to the case 63 by the third brake 68.
The third ring gear 42 can be connected to the output shaft 12 via the reduction gear pair 112 and the fourth clutch 114, and can also be connected to the output shaft 12 by a fifth clutch 116 provided in parallel with the fourth clutch 114. It is.

An output shaft gear 12a is provided integrally with the output shaft 12, and the output shaft gear 12a drives the wheels of the automobile through a not-shown counterpart gear.
Here, the reduction gear pair 112 is an output member of the present invention, has a role corresponding to the first reaction force member in the first embodiment, and the fourth clutch 114 serves as the fifth friction element of the present invention. The clutch 116 constitutes the second friction element of the present invention.

Next, the operation of the planetary gear train of the second embodiment shown in FIG. 4 will be described with reference to the operation table shown in FIG.
The planetary gear train of the second embodiment can also change the forward 8 speed and the reverse 1 speed by switching the engagement of the friction elements according to the operation table shown in FIG.

Although the detailed description is omitted, the fourth clutch 114 in the second embodiment is the same as the first brake 64 in the first embodiment, and the fifth clutch 116 in the second embodiment is the same as the second brake 66 in the first embodiment. It is.
That is, in forward first speed to fifth speed and reverse, the fourth clutch 116 that transmits output torque from the speed reduction gear pair 112, which is the output member of the present invention, to the output shaft 12 has low speed and reverse output torque. This is the same in that a corresponding large torque acts and the fifth clutch 116 provided in parallel with it energizes the torque transmission of the fourth clutch 116.

  As described in the first embodiment, the output torque transmitted from the reduction gear pair 112 to the output shaft 12 has a magnitude corresponding to the gear ratio, and is larger as the speed is lower. For this reason, if only the fourth clutch 114 is to be transmitted, it is necessary to have a capacity for transmitting the output torque of the first forward speed, and also to transmit a small output torque of the fifth forward speed. The drag torque at the stage, the problem of fuel consumption deterioration and heat generation, and the problem of the shift shock in the shift from the high speed stage to the fifth speed are the same as those of the first brake 64 described in the first embodiment.

However, unlike the first embodiment, the fifth clutch 116 is not interlocked with other friction elements and is engaged with an independent hydraulic pressure, so that the gear stages to be engaged with the first brake 64 are the first speed, the second speed, The fifth clutch 116 is engaged with the first clutch 64 not only in reverse but also in the third speed.
Therefore, the fourth clutch 64 only needs to be able to transmit the output torque of the fourth speed, and accordingly, the drag torque at the high speed stage, the fuel consumption deterioration and the problem of heat generation associated therewith can be reduced.

  As described above, the multi-stage planetary gear train and the friction element thereof according to the second embodiment of the present invention can reduce the deterioration of fuel consumption and the temperature increase of the transmission by reducing the drag torque at the sixth speed or higher. There is a merit that the control is easy to perform because the shift shock is less likely to occur in the shift from the high speed to the fifth speed.

  As described above, the multi-stage planetary gear train and its friction element according to each embodiment of the present invention can provide an automatic transmission that is excellent in fuel efficiency, generates less heat, and has less shift shock in a forward 8-speed transmission. It becomes possible.

  Since it is possible to obtain an automatic transmission that is excellent in fuel efficiency and generates little heat, it can be widely applied to medium-sized passenger cars to large commercial vehicles where fuel efficiency is important.

FIG. 2 is a skeleton diagram showing a multi-speed planetary gear train and its friction elements according to the present invention. (Example 1) It is a figure which shows the action | operation table | surface of the friction element of Example 1. FIG. 1 is a cross-sectional view of a main part of Example 1. FIG. FIG. 2 is a skeleton diagram showing a multi-speed planetary gear train and its friction elements according to the present invention. (Example 2) It is a figure which shows the action | operation table | surface of the friction element of Example 2. FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine 2 Torque converter 10 Input shaft 12 Output shaft 14 1st planetary gear set 16 2nd planetary gear set 18 3rd planetary gear set 19 4th planetary gear set 20 1st sun gear 22 1st ring gear 24 1st pinion 28 1st 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 Fourth Pinion 58 Fourth carrier 60 First clutch 62 Second clutch 63 Case 64 First brake 66 Second brake 68 Third brake 70 Third clutch 72 First piston 74, 76 Friction plate 80 Belleville spring 82 Connecting member 84 Inner cone 90 Second Sunda 92 second piston 94, 96 friction plate 102 connecting member 104 disc spring 106 the outer cone 110 engaged gear pair 112 reduction gear pair 114 4th clutch 116, fifth clutch

Claims (3)

The first reaction force member of the multi-speed planetary gear train that can achieve eight forward speeds arranged between the input shaft and the output shaft can be fixed to the stationary part, or the output member and the output shaft can be connected, A first friction element engaged in forward first speed to fifth speed and reverse;
A second friction element arranged in parallel with the first friction element,
A friction element of a multi-speed planetary gear train characterized in that both the first friction element and the second friction element are fastened at least in forward first speed, second speed and reverse.   The first brake constituting the first friction element, the second brake constituting the second friction element, and the second reaction force member fixed to the stationary portion in the forward first speed, the second speed, and the reverse speed 2. The second brake according to claim 1, wherein the second brake is engaged in conjunction with the connection when both the first brake and the third brake are engaged. Multi-speed planetary gear train.   The multi-speed planetary gear train according to claim 1 or 2, wherein the second friction element uses conical friction.

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