JP2005172123A - Multi-stage shift planetary gear series - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To make an interstage ratio of reduction stages of nearly direct connection small in a multi-stage shift planetary gear series with large gear change ratio. <P>SOLUTION: The multi-stage shift planetary gear series has a first gear group 14 and a second planet gear group 22 and the first gear group 14 is capable of providing a plurality of change gear ratio including speed up with normal rotation between an input shaft 10 and a middle output member. The second planet gear group 22 is equipped with the middle input member connected with a middle output member, a low speed stage fixing member fixable to a case 88, a high speed stage input member connectable with the input shaft 10 and an output member connected with an output shaft 12 as rotation members. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

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

As planetary gear trains used for vehicle automatic transmissions, gears capable of multi-speed shifting with seven forward speeds have been put into practical use, with the main objective of improving vehicle fuel efficiency, exhaust characteristics, and acceleration performance.
A conventional 7-speed forward planetary gear train in practical use has a gear ratio of 7 forward speeds by 4 planet gears and 6 or 7 friction elements. (See Patent Document 1)

  However, when the above conventional planetary gear train is applied to a vehicle such as a large truck having a large total vehicle weight, in order to ensure acceleration performance and climbing performance and satisfy demands such as exhaust performance and fuel efficiency, Insufficient gear ratio ratio (maximum and minimum ratio of gear ratio), and especially reduce the inter-stage ratio (ratio between adjacent gear ratios) between gear ratios close to gear ratio 1 (direct connection). However, it is difficult to select and select a fine gear ratio according to the driving conditions.

That is, in order to secure acceleration performance and climbing performance and meet requirements such as exhaust performance and fuel efficiency when applied to vehicles such as heavy trucks with large total vehicle weight, the gear ratio considers start acceleration and climbing ability A large gear ratio and a small gear ratio such as a speed increase suitable for high-speed driving are required. A heavier vehicle weight requires a larger gear ratio width.
In addition, the greater the gear ratio width and the greater the number of gear stages, the greater the degree of freedom in selecting the gear ratio. Therefore, the gear ratio width, the gear stage number, and the gear ratio are important in order to satisfy fuel consumption and exhaust performance. .

Nevertheless, in the conventional planetary gear train, among the gear stages that are frequently used, the speed ratio in the speed-up gear stage can take a smaller value than in the past, but the speed reduction gear is close to the gear ratio 1. In general, the interstage ratio is limited to about 1.35 because of the ratio of the number of teeth of the planetary gear, and it is difficult to reduce the interstage ratio.
On the other hand, in a so-called parallel shaft type manual transmission or a parallel shaft type transmission in which the shift operation is automated, 16 forward speeds are put into practical use, and a gear ratio width exceeding 10 is secured and a gear ratio of 1 The gear ratio at the speed reduction stage close to the direct connection is set to around 1.2. However, even in an automatic transmission using a planetary gear train, the speed reduction near the direct connection is achieved with the large speed ratio width and the number of speed stages as described above. It is desirable to reduce the interstage ratio in the stages.
JP 2000-266138 A

The problems to be solved are that the gear ratio width is small and the number of gears is small, and it is difficult to reduce the gear ratio at the speed reduction gear close to the direct connection. When applied to automatic transmissions, the gear ratio can be selected with a low degree of freedom, ensuring acceleration performance and climbing performance, but not meeting fuel efficiency and exhaust performance requirements. Is a difficult point.
The object of the present invention is applicable to an automatic transmission such as a large truck having a large total vehicle weight, and has a large gear ratio range, a large number of gears, and a reduction in gear ratio at a reduction gear near direct coupling. Is to provide a multi-speed planetary gear train.

The most important aspect of the present invention is that, in the first gear group on the upstream side (the side closer to the engine), a speed ratio of three to five forward rotations capable of increasing speed in addition to at least a plurality of speed reductions and direct couplings can be obtained. Features.
That is, the multistage planetary gear train of the present invention is provided between the input shaft and the output shaft, and converts the rotation speed of the input shaft into the rotation speed of the output shaft and the first gear group on the upstream side and the second gear on the downstream side. The first gear group has an intermediate output member that inputs the output of the first gear group to the second planetary gear group as a rotating member, and is positive between the input shaft and the intermediate output member. It is configured to obtain a plurality of speed ratios including speed increase in the rolling direction, the second gear group is composed of two planetary gear sets, and as a rotating member, an intermediate input member that can be connected to or connected to the intermediate output member, A low-speed stage fixed member that can be fixed to the case, a high-speed stage input member that can be connected to the input shaft, and an output member that is connected to the output shaft. In the common speed diagram of the second planetary gear group, High-speed input member, output member Each speed shaft of the low speed stage stationary members were arranged in this order.

  The multi-speed planetary gear train of the present invention has an intermediate output member that inputs the output of the first gear group to the second planetary gear group as a rotating member in the first gear group constituted by two upstream rows, and an input shaft And the intermediate output member are configured to obtain a plurality of speed ratios including speed increase in the forward rotation direction, so that the speed ratio ratio width is increased while being multistaged by the action of the second planetary gear group composed of two downstream rows. In addition to being widespread, the gear ratio of the deceleration stage, which is close to the direct connection, which is frequently used, can be reduced, so that it can satisfy acceleration performance and climbing performance when applied to heavy trucks with heavy vehicle weight. In addition, the exhaust characteristics and fuel consumption can be improved.

  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 each example.

FIG. 1 is a skeleton diagram of one embodiment of the device of the present invention, and shows the upper half of the concentric input shaft 10 and output shaft 12 from the axis.
In the multi-stage planetary gear train of the present invention shown in FIG. 1, the input shaft 10 and the output shaft 12 are coaxially arranged, and three sets on the upstream side (left side in the figure) on these shafts are the first gear group. 14, and is composed of a first planetary gear set 16, a second planetary gear set 18, and a third planetary gear set 20.
Further, two sets on the downstream side (right side in the figure) are the second planetary gear group 22, which is composed of a fourth planetary gear set 24 and a fifth planetary gear set 26.

The first planetary gear set to the fifth planetary gear set 16, 18, 20, 24, 26 are all generally called single pinion types, and have the same configuration.
That is, the first planetary gear set 16 includes a first sun gear 30, a first ring gear 32, a first pinion 34 that meshes with the first ring gear 32 and the first sun gear 30, and a first pinion 34 that is rotatably supported. And the first carrier 38.

  Similarly, the second planetary gear set 18 is the second sun gear 40, the second ring gear 42, the second pinion 44, and the second carrier 48, and the third planetary gear set 20 is the third sun gear 50, the third ring gear 52, In the third pinion 54 and the third carrier 58, the fourth planetary gear set 24 is in the fourth sun gear 60, the fourth ring gear 62, the fourth pinion 64 and the fourth carrier 68, and in the fifth planetary gear set 26 is the fifth planetary gear set 26. The sun gear 70, the fifth ring gear 72, the fifth pinion 74, and the fifth carrier 78 are respectively configured.

The input shaft 10, the output shaft 12, and the rotating members of the first to fifth planetary gear sets 16, 18, 20, 24, and 26 are connected as follows or can be connected.
The input shaft 10 is connected to the second sun gear 40, and is connected to the first carrier 38 and the second ring gear 42 connected to each other via the first clutch 80 and to the fourth carrier connected to each other via the second clutch 82. The ring gear 62 and the fifth carrier 78 can be selectively connected to the first sun gear 30 via the third clutch 84, respectively.

The first carrier 38 and the second ring gear 42 can be fixed to the case (stationary part) 88 by the first brake 86, and one rotation direction is always fixed to the case 88 by the one-way clutch 90.
The first sun gear 30 can be fixed to the case 88 by the second brake 92.
The first ring gear 32 is connected to the third carrier 58 and to the fifth ring gear 72.

The second carrier 48 and the third ring gear 52 are connected.
The third sun gear 50 can be fixed to the case 88 by the third brake 94.
The fourth sun gear 60 is connected to the fifth sun gear 70 and can be fixed to the case 88 by a fourth brake 96.
The fourth carrier 68 is connected to the output shaft 12.

Here, the third carrier 58 and the first ring gear 32 in the first gear group 14 and connected to the fifth ring gear 72 of the second planetary gear group 14 constitute an intermediate output member of the present invention.
Further, the fifth ring gear 72 in the second planetary gear group 22 and connected to the intermediate output member of the first gear group 14 constitutes the intermediate input member of the present invention.
The fourth ring gear 62 and the fifth carrier 78 in the second planetary gear group 22 and connectable to the input shaft 10 constitute a high-speed stage input member of the present invention.
And the 4th sun gear 60 and the 5th sun gear 70 which are in the 2nd planetary gear group 22 and can be fixed to case 88 constitute the low speed stage fixed member of the present invention.
Further, the fourth carrier 68 in the second planetary gear group 22 and connected to the output shaft 12 constitutes the output member of the present invention.

Next, the operation of the first embodiment shown in FIG. 1 will be described with reference to the operation table shown in FIG. 2 and the common velocity diagrams shown in FIGS.
FIG. 3 is a common speed diagram for the first gear group 14, and FIG. 4 is a common speed diagram for the second planetary gear group 22.
In the following description, the clutch and the brake are referred to as a friction element, and the one including the one-way clutch and having a connecting function between the shaft, the stationary portion, and the rotating member is collectively referred to as a fastening element.

In the operation table of FIG. 2, fastening elements such as a clutch, a brake, and a one-way clutch are assigned to the horizontal column, C-1 represents the first clutch 80, B-1 represents the first brake 86, OC represents the one-way clutch 90 and so on.
The relationship between these symbols and the symbols of the respective fastening elements is shown in FIG.

The vertical column is divided into “D range”, “R range” and “L range” of the control lever (not shown). The D range is the first forward speed (1st) to the 14th speed (14th), and the R range is Each reverse speed (Rev) is assigned.
In the L range, the input shaft 10 side can be driven from the output shaft 12 side like an engine brake described later.
In the operation table of FIG. 2, “o” indicates the fastening of each fastening element, and the blank indicates the release of each fastening element.
In addition, (◯) mark indicates that it is fastened but not involved in power transmission.

  The common speed diagrams shown in FIGS. 3 and 4 represent the rotational speed of each rotating member when the rotational speed (rotational speed) of the input shaft 10 is 1 in the vertical direction, and the first planetary gear set through the horizontal direction. The rotation members are assigned to intervals corresponding to the respective gear ratios of the fifth planetary gear sets 16, 18, 22, 24, and 26, and the speed axis is drawn with a vertical line for each rotation member.

  Symbols surrounded by a line above each speed axis in the common speed diagram are S for the sun gear, R for the ring gear, C for the carrier, and the following numbers 1 and 2 to which the first to fifth planets belong respectively. For example, S1, R1, and C1 represent the first sun gear 30, the first ring gear 32, and the first carrier 38 of the first planetary gear set 16, respectively.

In the common speed diagram, the height of the intersection of the vertical line (speed axis) of each rotating member and the thick line represents the number of rotations of each rotating member, and the horizontal thin alternate long and short dash line indicates the intermediate output member and intermediate input member. And are connected at the same rotational speed.
The numbers and Rev enclosed in <> are numbers for the gear ratio in the first gear group 14 to be described later.
In FIG. 4, for easy understanding, the intersections on the vertical lines of the fourth carrier 68 (C4) connected to the output shaft 12 are indicated by small ◯ marks.
In addition, the intersection points on the vertical lines of the first ring gear 32 (R1) and the third carrier 58 (C3) of the intermediate output member are also indicated by Δ.
Furthermore, the portion where the intersection is represented by ● indicates the arrangement of the fastening elements.

Here, the gear ratio of each planetary gear set is 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 16 is α1, the second planetary gear set. The gear set 18 is α2, the third planetary gear set 20 is α3, the fourth planetary gear set 24 is α4, and the fifth planetary gear set 26 is α5.
In the calculation of the gear ratio including the common speed diagram, the case where α1 is 0.35, α2 is 0.43, α3 is 0.37, α4 is 0.47, and α5 is 0.60 is explained. To do.
Further, forward rotation indicates the same rotation direction as the input shaft 10, and reverse rotation indicates the opposite rotation direction.

First, the gear ratio (the rotational speed of the input shaft 10 / the rotational speeds of the third carrier 58 and the first ring gear 32) in the first gear group 14 will be described.
The first speed ratio i1 of the forward rotation is obtained by fixing the third sun gear 50 to the case 88 by fastening the third brake 94.
At this time, since the second ring gear 42 is always fixed to the case 88 by the action of the one-way clutch (OC) 90 in the direction of driving (accelerating) the vehicle, the first speed ratio i1 of normal rotation is obtained.

When this is represented in a common velocity diagram, it becomes as shown in FIG. 3A, and the speed change is performed by the second planetary gear set 18 and the third planetary gear set 20. That is, the rotation of the second sun gear (S2) 40 is the same as that of the input shaft 10, and the rotation speeds of the second ring gear (R2) 42 and the third sun gear (S3) 50 are hatched.
The intersection (Δ mark) between this oblique line and the vertical line of the third carrier (C3) 58 is the rotational speed of the third carrier 58 indicated by <1>.
The first speed ratio i1 for forward rotation is {(1 + α2) (1 + α3)} / α2, and is 4.556 in the above-described gear ratio.

  As described above, when the vehicle is driven, the second ring gear 42 is fixed by the action of the one-way clutch 90. In this case, however, the input shaft 10 cannot be driven from the output shaft 12 side like an engine brake. Therefore, when driving from the output shaft 12 side, the second ring gear 42 is fixed to the case 88 by fastening the first brake 86.

Next, the second speed ratio i2 of the forward rotation is such that the first sun gear 30 is fixed to the case 88 by fastening the second brake 92 while the third sun gear 50 is fixed to the case 88 by the third brake 94 as it is. It is obtained by.
At this time, the fixing of the second ring gear 42 to the case 88 is automatically released by the action of the one-way clutch 90.
The common speed diagram is as shown in FIG. 3B, and the rotation speeds of the first ring gear 32 and the third carrier 58 are as shown by <2>.
The second speed ratio i2 for forward rotation is {α2 (1 + α1) + α3 (1 + α1) (1 + α2) + α1} / {α2 (1 + α1)}, which is 2.833 in the above-described gear ratio.

Next, the forward third rotation speed ratio i3 is determined by releasing the engagement of the second brake 92 and engaging the third clutch 84 while the third sun gear 50 is fixed to the case 88 by the third brake 94. It is obtained by.
That is, as shown in the common speed diagram shown in FIG. 3C, the first sun gear 30 and the second sun gear 40 have the same rotational speed as the input shaft 10, and the rotational speed of the intermediate output member is <3>. Become.
As a result, the third speed ratio i3 for forward rotation is {α3 (1 + α2) + α1 (1 + α2) (1 + α3) + α2} / {α2 + α1 (1 + α2)}, which is 1.768 in the above-described gear ratio.

Next, the forward fourth speed ratio i4 is determined by releasing the engagement of the third clutch 84 and engaging the first clutch 80, while the third sun gear 50 is fixed to the case 88 by the third brake 94. It is obtained by integrating the two planetary gear set 18 with the input shaft 10.
That is, as shown in the common speed diagram shown in FIG. 3D, the third ring gear 52 has the same rotational speed as the input shaft 10, and the rotational speed of the intermediate output member is <4>.
As a result, the forward fourth speed ratio i4 is 1 + α3, which is 1.370 in the above-described gear ratio.

Next, in the forward fifth speed ratio i5, the third clutch 94 is released from the fixing of the third sun gear 50 to the case 88 and the third clutch 84 is engaged again with the first clutch 80 engaged. It is obtained by doing.
As a result, the entire first planetary gear group 14 is integrated and directly connected to the input shaft 10, and the fifth forward gear ratio i5 is 1 regardless of the gear ratio. This is indicated by <5> in FIG.

Next, the forward sixth rotation gear ratio i6 is obtained by fixing the first sun gear 30 to the case 88 again by engaging the second brake 92 with the first clutch 80 engaged.
3 (e), the first carrier 38 has the same rotational speed as the input shaft 10, the first sun gear 30 has the rotational speed 0, and the intermediate output member rotates. The number is <6>.
As a result, the forward sixth gear ratio i6 is determined only by the first planetary gear set 16, becomes 1/1 + α1, and is 0.741 in the above-mentioned gear ratio.

Next, the reverse reduction ratio ir is obtained by connecting the first sun gear 30 to the input shaft 10 by engaging the third clutch 84 and fixing the first carrier 38 to the case 88 by engaging the first brake 86. .
That is, as shown in the common velocity diagram shown in FIG. 3E, the speed change is performed only by the first planetary gear set 16, and the rotation speed of the intermediate output member is <Rev>.
As a result, the reverse reduction ratio ir is −1 / α1, and is −2.857 in the above-described number ratio of teeth.

Next, although it cannot be said to be a speed change, when the first and second brakes 86 and 92 are fastened and the first sun gear 30 and the first carrier 38 are fixed to the case 88, the first planetary gear set 16 is integrated. Even if the input shaft 10 is fixed, the rotation of the intermediate output member can be made zero.
In (e) of FIG. 3, it represents with <0>.

Next, the overall gear ratio will be described based on the gear ratios in the first gear group 14.
Here, in order to simplify the following display and calculation formula, the intermediate value A is determined as follows.
A = {α5 (1 + α4) + α4} / {α5 (1 + α4)}
Substituting the above gear ratio into A gives 1.533.

Next, the operation of each gear stage will be described with reference to the operation table of FIG. 2 and the common speed diagram of the second planetary gear group 22 shown in FIG.
Driving in the first forward speed (1st) is performed so that the first gear group 14 obtains the first gear ratio i1 of normal rotation, and the fourth brake 96 is engaged and the fourth and fifth sun gears 60 are engaged. , 70 is fixed to the case 88.
As a result, the fifth ring gear 72 of the intermediate input member is driven at the first speed ratio i1 of the forward rotation and becomes the rotation indicated by <1>, and the rotation of the fourth and fifth sun gears 60 and 70 becomes zero. In the common speed diagram of the second planetary gear group 22, the diagonal line indicated by 1 and the intersection with the vertical line of the fourth carrier (C 4) is the rotational speed of the output shaft 12.
The gear ratio is i1 (1 + α4) (1 + α5), and the above-mentioned gear ratio is 10.716.

Next, the forward shift to the second speed (2nd) is performed by setting the first gear group 14 to the forward second speed ratio i2 while maintaining the engagement of the fourth brake 96 at the first speed. .
As a result, the diagonal line indicated by 2 in the common speed diagram is shown, the gear ratio is i2 (1 + α4) (1 + α5), and the above-mentioned gear ratio is 6.664.

Next, the forward shift to the third speed (3rd) is performed by setting the first gear group 14 to the forward third speed ratio i3 while maintaining the engagement of the fourth brake 96 at the second speed. .
As a result, the oblique line indicated by 3 in the common speed diagram is obtained, the gear ratio is i3 (1 + α4) (1 + α5), and the above-mentioned gear ratio is 4.157.

Next, the forward shift to the fourth speed (4th) is performed by setting the first gear group 14 to the forward fourth speed ratio i4 while maintaining the engagement of the fourth brake 96 at the third speed. .
As a result, the diagonal line indicated by 4 in the common speed diagram is shown, the transmission ratio is i4 (1 + α4) (1 + α5), and the above-mentioned gear ratio is 3.222.

Next, in the forward shift to the fifth speed (5th), the first gear group 14 is set to the fifth speed ratio i5 (direct connection) of the forward rotation while maintaining the engagement of the fourth brake 96 at the fourth speed. Do that.
As a result, the oblique line indicated by 5 in the common speed diagram is obtained, the gear ratio is (1 + α4) (1 + α5), and the above-mentioned gear ratio is 2.352.

Next, the forward shift to the sixth speed (6th) is performed by setting the first gear group 14 to the forward sixth speed ratio i6 while maintaining the engagement of the fourth brake 96 at the fifth speed. .
As a result, the diagonal line indicated by 6 in the common speed diagram is obtained, the gear ratio is i6 (1 + α4) (1 + α5), and the above-mentioned gear ratio is 1.742.

Next, the forward shift to the seventh speed (7th) is performed by engaging the second clutch 82 while maintaining the engagement of the fourth brake 96 at the sixth speed, whereby the fourth ring gear 62 of the high speed input member is engaged. The fifth carrier 78 is connected to the input shaft 10.
As a result, the diagonal line indicated by 7 in the common speed diagram is obtained, the gear ratio is (1 + α4), and the above-mentioned gear ratio is 1.470.

Next, in the forward shift to the eighth speed (8th), the engagement of the fourth brake 96 is released while maintaining the engagement of the second clutch 82 at the seventh speed, and the first planetary gear group 14 is moved again. This is done by setting the sixth speed ratio i6 for forward rotation.
As a result, the diagonal line indicated by 8 in the common speed diagram is obtained, the gear ratio is i6 / {1-A (1-i6)}, and the above-mentioned gear ratio is 1.229.

Next, in the forward shift to the ninth speed (9th), the first planetary gear group 14 is again rotated in the forward fifth speed ratio i5 (direct connection) while maintaining the engagement of the second clutch 82 at the eighth speed. To do.
As a result, the entire first and second planetary gear groups 14 and 16 are integrated into a horizontal line indicated by 9 in the common speed diagram, and the gear ratio is 1 (direct connection) regardless of the gear ratio. .

Next, in the forward shift to the 10th speed (10th), the first planetary gear group 14 is again set to the forward fourth speed ratio i4 while maintaining the engagement of the second clutch 82 at the ninth speed. To do.
As a result, the oblique line indicated by 10 in the common speed diagram is obtained, the gear ratio becomes i4 / {1 + A (i4-1)}, and the speed increase is 0.874 in the above-described gear ratio.

Next, in the forward shift to the 11th speed (11th), the first planetary gear group 14 is again set to the third speed ratio i3 of the forward rotation while maintaining the engagement of the second clutch 82 at the 10th speed. To do.
As a result, the diagonal line indicated by 11 in the common speed diagram is obtained, the gear ratio is i3 / {1 + A (i3-1)}, and the above-mentioned gear ratio is 0.812.

Next, in the forward shift to the 12th speed (12th), the first planetary gear group 14 is again set to the second speed ratio i2 of the forward rotation while maintaining the engagement of the second clutch 82 at the 11th speed. To do.
As a result, the oblique line indicated by 12 in the common speed diagram is obtained, and the gear ratio is i2 / {1 + A (i2-1)}, which is 0.744 in the above-mentioned gear ratio.

Next, in the forward shift to the 13th speed (13th), the first planetary gear group 14 is again set to the first speed ratio i1 of the forward rotation while maintaining the engagement of the second clutch 82 at the 12th speed. To do.
As a result, the diagonal line indicated by 13 in the common speed diagram is obtained, and the gear ratio becomes i1 / {1 + A (i1-1)}, which is 0.706 in the above-mentioned gear ratio.

Next, in the forward shift to the 14th speed (14th), the first planetary gear set is engaged by engaging the first and second brakes 86 and 92 while maintaining the engagement of the second clutch 82 at the 13th speed. 16 is integrated into <0>, that is, fixed to the case 88.
As a result, the diagonal line indicated by 14 in the common speed diagram is obtained, and the gear ratio is {α5 (1 + α4)} / {α4 + α5 (1 + α4)}, which is 0.652 in the above-described gear ratio.

Next, in the reverse (Rev) drive in the R range, the first gear group 14 is made to obtain the reverse speed ratio ir, and the fourth brake 96 is engaged and the fourth and fifth sun gears 60 and 70 are engaged. This is done by fixing to the case 88.
As a result, the hatched line indicated by Rev in the common speed diagram is obtained, and the gear ratio is ir (1 + α4) (1 + α5), which is −6.720 in the above-described gear ratio.
That is, in reverse, the torque reversely decelerated by the first gear group 14 is further decelerated and driven by the second planetary gear group 22.

As described above, since the one-way clutch 90 is automatically engaged only in the direction of accelerating the vehicle at the first speed in the D range, when driving from the output shaft 12 side like an engine brake, FIG. As indicated by 1st in the L range of the operation table, the first brake 86 is engaged in addition to the engagement of the third and fourth brakes 94 and 96.
Thereby, the gear ratio of the forward first speed can be obtained regardless of the direction in which the torque acts.
Since there is the one-way clutch 90, it is possible to easily perform control for suppressing the occurrence of a so-called shift shock in the shift from the first speed to the second speed.
That is, the occurrence of a shift shock can be suppressed only by gradually engaging the second brake 92.

The above gear ratio is summarized as follows. Note that the ratio between adjacent gear ratios is the inter-step ratio and is shown in parentheses. That is, the value shown for the first speed is a ratio with the gear ratio of the second speed.
1st speed 10.716 (1.608)
2nd speed 6.664 (1.603)
3rd speed 4.157 (1.290)
4th speed 3.222 (1.370)
5th speed 2.352 (1.350)
6th speed 1.742 (1.185)
7th speed 1.470 (1.196)
8th speed 1.229 (1.229)
9th speed 1.000 (1.144)
10th speed 0.874 (1.076)
11th speed 0.812 (1.091)
12th speed 0.744 (1.054)
13th speed 0.706 (1.083)
14th speed 0.652

  As can be seen, the gear ratio range (10.716 / 0.652 = 16.436) is wide and the number of gears is large, and the gear ratio generally decreases with increasing speed. Therefore, the transmission ratio of the vehicle transmission driven by the internal combustion engine tends to be favorable.

It is particularly characterized by the fact that the speed ratio is close to the direct connection that has been frequently used and has been difficult to achieve a small interstage ratio, that is, the interstage ratio from the sixth speed to the ninth speed is about 1.2. It is.
The reason why the interstage ratio in the speed reduction stage close to the direct connection can be reduced is that the speed increase stage exists in the first planetary gear group 14 as can be seen from the common speed diagram of FIG.

That is, when the rotational speed of the intermediate input member of the second planetary gear group 22 is driven by direct coupling or speed increase in the first planetary gear group 14, the gear ratio from the sixth speed to the ninth speed is obtained. It is done.
For this reason, when applied to an automatic transmission such as a large truck with a large gross vehicle weight, the gear ratio can be selected freely, and the acceleration performance can be improved by selecting the optimum gear ratio according to the driving conditions. There are advantages in improving fuel efficiency.

FIG. 5 is a skeleton of the second embodiment in the multi-speed planetary gear train 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 first one-way clutch (OC1) 90 is disposed so as to fix the first sun gear 30 to the case 88, and the third sun gear 50 is connected to the second one-way clutch. (OC2) 98 and the fifth brake 100 enable fixing to the case 88 only in one rotation direction.
As a result, the first planetary gear group 14 can be shifted except for the first forward gear ratio in the first embodiment.

Further, since the second one-way clutch (OC2) 98 and the fifth brake 100 are added, the control for suppressing the shift shock becomes easy.
As can be seen from the operation table shown in FIG. 6, the first one-way clutch 90 is effective in shifting from the first speed (1st) to the second speed (2nd) in the D range. Since only the fastening is performed, the shift shock can be suppressed only by controlling so that the fastening is performed gradually.

In the second embodiment, it is possible to easily control the shift from the second speed to the third speed and the shift from the third speed to the fourth speed.
That is, in the shift from the second speed to the third speed, the engagement of the third clutch 84 is released and the first clutch 80 is engaged. At this time, both clutches 84 and 80 transmit torque simultaneously. Even if there is, there is only a temporary shift to the next fourth speed, and there is no shock in the direction of decelerating the vehicle.

Therefore, the control in the shift from the second speed to the third speed is also facilitated, and in the next shift to the fourth speed, it is only necessary to add the engagement of the third clutch 84 again. Since it is possible to suppress the shift shock only by performing the control so as to be performed, it can be performed easily.
That is, by adding the second one-way clutch (OC2) 98 and the fifth brake 100, in addition to the first speed to the second speed, the shift from the second speed to the third speed and the third speed to the fourth speed. It is possible to make it possible to easily control the gear shift.

  When the first one-way clutch 90 and the second one-way clutch 98 are used, driving from the output shaft 12 cannot be performed, but driving from the output shaft 12 by engine braking or the like at the first to third speeds. To make this possible, it is combined with the engagement of the second brake 92 and the third brake 94 as shown in the L range of the operation table.

On the other hand, as described above, since there is no first gear ratio i1 in the first planetary gear group 14 as described above, the overall gear ratio is accordingly changed from the first speed of the forward speed to the thirteenth speed from the example of the first embodiment. It becomes the gear ratio excluding the speed.
Although detailed description is omitted, when the gear ratio is the same as the value described in the first embodiment, each gear ratio is as follows. The values in parentheses are the step-to-step ratios between the adjacent gear ratios, and the brackets << in the high-speed gears after the fifth speed are the step-to-step ratios when shifting by one step.
1st speed 6.664 (1.603)
Second gear 4.157 (1.290)
3rd speed 3.222 (1.370)
4th speed 2.352 (1.350)
5th speed 1.742 (1.185) << 1.417 >>
6th speed 1.470 (1.196) << 1.470 >>
7th speed 1.229 (1.229) << 1.406 >>
8th speed 1.000 (1.144) << 1.232 >>
9th speed 0.874 (1.076) << 1.175 >>
10th speed 0.812 (1.092) << 1.245 >>
11th speed 0.744 (1.141)
12th speed 0.652
Reverse -6.720

As can be seen, a transmission gear ratio of 12 forward speeds is obtained, the ratio of the transmission gear ratio of the first speed to the 12th speed is a large value of 10.581, and the interstage ratio also decreases as the speed increases. Therefore, the gear ratio of the vehicle transmission driven by the internal combustion engine tends to be favorable, and the gear ratio close to the direct connection, which is particularly frequently used, that is, the gear ratio from the fifth speed to the eighth speed is implemented. Similar to Example 1, it is characterized by being about 1.2.
The reason why the interstage ratio can be reduced is as described in the first embodiment.

  For this reason, when applied to an automatic transmission such as a large truck with a large gross vehicle weight, the gear ratio can be selected freely, and the acceleration performance can be improved by selecting the optimum gear ratio according to the driving conditions. There are advantages in improving fuel efficiency.

Further, the speed ratio of the forward first speed and the speed ratio of the reverse drive are substantially the same value (absolute value).
For this reason, even when a friction clutch generally used in a manual transmission vehicle is used without using a torque converter between an internal combustion engine (not shown) and the input shaft 10, a similar driving force can be obtained in forward and reverse travel. Therefore, it is possible to apply the friction clutch to lower the manufacturing cost and further improve the fuel consumption.

  Accordingly, the second embodiment can be easily applied to a large truck having a heavy vehicle weight, and can be easily configured without using a torque converter, which is advantageous in improving manufacturing cost and fuel consumption.

FIG. 7 is a skeleton of the third embodiment in the multi-speed planetary gear train 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 configuration of the first gear group 14 is different.
That is, the first gear group 14 is composed of two sets of a first planetary gear set 16 and a second planetary gear set 18.

The connection relationship of the rotating members around the first gear group 14 is as follows.
The input shaft 10 includes the second sun gear 40 via the first clutch 80, the fourth ring gear 62 and the fifth carrier 78 connected to each other via the second clutch 82, and the first sun gear 30 via the third clutch 84. The first carrier 38 and the second ring gear 42 that are connected to each other via the fourth clutch 102 can be selectively connected to each other.
The first carrier 38 and the second ring gear 42 can be fixed to the case 88 by a first brake 86 and are always fixed to the case 88 in one rotational direction by a one-way clutch (OC) 90.
The first sun gear 30 can be fixed to the case 88 by the second brake 92.
Since the connection relationship in the second planetary gear group 22 is the same as that in the first embodiment, the description thereof is omitted.

Next, the operation of the third embodiment will be described.
The gear ratio of each planetary gear set 16, 18, 24, 26 is exemplified for α1 of 0.42, α2 of 0.62, α4 of 0.45, and α5 of 0.60.
The first planetary gear group 14 is a gear train generally used in an automatic transmission having four forward speeds and one reverse speed, and each gear ratio can be obtained as follows.
The forward first speed ratio i <b> 1 is obtained by connecting the input shaft 10 and the second sun gear 40 by engaging the first clutch 80. At this time, the first carrier 38 and the second ring gear 42 are fixed to the case 88 in the direction of driving the vehicle by the one-way clutch 90.

As shown in the common speed diagram of the first planetary gear group 14 shown on the left side of FIG. 8, the second sun gear 40 has the same rotational speed 1 as the input shaft 10, and the second ring gear 42 has a rotational speed of 1,8. The gear ratio i1 is (1 + α2) / α2, and the above-mentioned gear ratio is 2.613.
Note that “1, 8” written in the diagonal line of the common speed diagram means that the rotation at the first speed and the eighth speed is shown as will be described later. It is.

  Further, when the forward first speed ratio is obtained by the one-way clutch 90, driving from the output shaft 12 side is not possible, but when the first brake 86 is engaged and the second ring gear 42 is fixed to the case 88. As in the first embodiment, the gear ratio can be obtained regardless of the direction in which the torque acts.

Next, the forward second speed ratio i2 is obtained by fixing the first sun gear 30 to the case 88 by fastening the second brake 92 while keeping the first clutch 80 engaged.
As a result, as shown in the common speed diagram, the first sun gear 30 has a diagonal line written as 2 and 7, where the rotation speed is 0, and the speed ratio i2 is {α2 (1 + α1) + α1} / {α2 (1 + α1)}. Thus, the above-described ratio of the number of teeth is 1.477.

Next, the forward third rotation speed ratio i3 is determined by releasing the engagement of the second brake 92 while maintaining the engagement of the first clutch 80 and engaging the first clutch 38 and the second ring gear 42 when the fourth clutch 102 is engaged. Is obtained by connecting to the input shaft 10.
As a result, the entire first planetary gear group 14 is integrated and directly connected to the input shaft 10 as seen in the horizontal line written as 3 and 6 shown in the common velocity diagram.
The third gear ratio i3 is 1 regardless of the gear ratio.

Next, the forward fourth rotation speed ratio i4 is determined by releasing the engagement of the first clutch 80 without changing the engagement of the fourth clutch 102, and re-engaging the second brake 92 to move the first sun gear 30 to the case 88. Obtained by fixing.
As a result, as shown in the common speed diagram, the first sun gear 30 has a diagonal line written as 4, 5, where the rotational speed is 0, and the gear ratio i4 becomes 1 / (1 + α1), and the above-described gear ratio In this case, the speed increases to 0.704.

Next, the reverse speed ratio ir is determined by connecting the first sun gear 30 and the input shaft 10 by engaging the third clutch 84 and fixing the first carrier 38 to the case 88 by engaging the first brake 86. It is obtained by.
As a result, it is seen in the oblique line written as Rev in the common speed diagram, and the gear ratio ir is −1 / α1, and in the above-mentioned gear ratio, it is −2.381.

Next, the overall gear ratio will be described based on the gear ratios in the first gear group 14.
Here, the intermediate value A for simplifying the following display and calculation formula is the same calculation as in Example 1, and is 1.517 when the above-mentioned tooth ratio is used.

Next, the operation of each gear stage will be described with reference to the operation table of FIG. 7 and the common speed diagram of the second planetary gear group 22 shown on the right side of FIG.
Since the operation of the second planetary gear group 22 is the same as that of the first embodiment, the fourth and fifth sun gears 60 and 70 of the low speed stage fixed members are fixed to the case 88 from the first forward speed to the fourth speed. To change the speed.

  That is, when the first planetary gear group 14 is set to the first speed ratio i1, the forward first speed ratio is obtained, and the speed ratio is i1 (1 + α4) (1 + α5), and when the gear ratio is set as above, It becomes 6.062.

  Further, when the first planetary gear group 14 is set to the second speed ratio i2, the forward second speed gear ratio is obtained, and the speed ratio is i2 (1 + α4) (1 + α5). It becomes 3.427.

  Next, when the first planetary gear group 14 is set to the third speed ratio i3, the speed ratio of the third forward speed is obtained, and the speed ratio is (1 + α4) (1 + α5), and when the above gear ratio is set, It becomes 2.320.

  Subsequently, when the first planetary gear group 14 is set to the fourth gear ratio i4, the gear ratio of the forward fourth speed is obtained, and the gear ratio is i4 (1 + α4) (1 + α5), and the above gear ratio is used. 1.634.

Next, at the high speed stage after the fifth speed, the fourth ring gear 62 and the fifth carrier 78 of the high speed stage input member are connected to the input shaft 10 to obtain each gear ratio.
That is, when the first planetary gear group 14 is set to the fourth speed ratio i4, the forward fifth speed ratio is obtained, the speed ratio is i4 / {1-A (1-i4)}, and the number of teeth The ratio is 1.278.

Further, when the first planetary gear group 14 is set to the fourth gear ratio i3, a forward sixth speed gear ratio is obtained, and the gear ratio is directly connected to 1 regardless of the gear ratio.
Next, when the first planetary gear group 14 is set to the second gear ratio i2, the forward seventh gear ratio is obtained, and the gear ratio is i2 / {1 + A (i2-1)}. In this case, the speed is increased to 0.857.

  Subsequently, when the first planetary gear group 14 is set to the first speed ratio i1, the speed ratio of the eighth forward speed is obtained, and the speed ratio is i1 / {1 + A (i1-1)}. Then, it becomes 0.758.

Further, when the first brake 86 and the second brake 92 are engaged and the first planetary gear set 16 is integrally fixed to the case 88, the first ring gear 32 as the intermediate output member and the fifth ring gear 72 as the intermediate input member. Becomes 0, and the gear ratio of forward ninth speed is obtained.
The gear ratio of the ninth speed is {α5 (1 + α4)} / {α4 + α5 (1 + α4)}, which is 0.659 in the above-described gear ratio.

Subsequently, the reverse drive is performed by fixing the fourth and fifth sun gears 60 and 70 of the low speed stage fixed member to the case 88 and setting the first planetary gear group 14 to the reverse speed ratio ir.
The reverse gear ratio is ir (1 + α4) (1 + α5), and the above-mentioned gear ratio is −5.524.

These gear ratios are summarized as follows. The values in parentheses are the inter-step ratios between adjacent gear ratios.
1st Speed 6.062 (1.769)
Second gear 3.427 (1.477)
3rd speed 2.320 (1.420)
4th speed 1.634 (1.279)
5th speed 1.278 (1.278)
6th speed 1.000 (1.167)
7th speed 0.857 (1.130)
8th speed 0.758 (1.150)
9th speed 0.659
Reverse -5.524

  As described above, in the third embodiment, it is possible to perform a shift of 9 forward speeds and 1 reverse speed, the ratio of the speed ratio between the first speed and the ninth speed is a large value of 9.197, and the interstage ratio is also high speed. As the gear goes to the stage, it becomes smaller and tends to be preferable as the gear ratio of the vehicle transmission driven by the internal combustion engine.

In particular, the speed ratio close to the direct connection, that is, the inter-step ratio from the fourth speed to the sixth speed is the same as in the first embodiment, and is characterized by a small value of 1.3 or less.
The reason why the interstage ratio can be reduced is the same as described in the first embodiment.
For this reason, when applied to an automatic transmission such as a truck having a large gross vehicle weight, the gear ratio can be freely selected. There is an advantage in improving fuel consumption.

  Although illustration of details is omitted, a second one-way clutch and a fifth clutch can be added in parallel with the first clutch 80, and the input shaft 10 and the second sun gear 40 can be connected by the second one-way clutch. Thus, the fifth clutch is always engaged during forward travel, so that the control for reducing the shift shock in the shift from the first speed to the fourth speed can be facilitated as in the second embodiment. it can.

Further, the reverse gear ratio can be obtained by another method for the reverse.
That is, although illustration is omitted, the third clutch 84 is deleted, and a brake for fixing the fourth ring gear 62 and the fifth carrier 78 to the case 88 is newly provided.
During reverse travel, the fourth ring gear 62 and the fifth carrier 78 are fixed to the case 88, and the first planetary gear group 14 is set to the forward first speed ratio i1 to obtain the reverse speed ratio.
In this case, the gear ratio is − {i1 · α5 (1 + α4)} / α4.
In this case as well, there is no change in the forward gear ratio, so the characteristics of the third embodiment described above remain unchanged.

  Although the first to third embodiments have been described above, what is common to these is that in the first gear group 14 constituted by two or three rows upstream of the planetary gears of four or five rows, The most important feature is to use a gear train having a structure that can obtain a gear ratio of at least four forward rotations including speed increase.

As a result, it is possible to obtain a gear ratio of 9 or more forward speeds with a large speed ratio width and a preferable gear ratio.
In particular, when used in an automatic transmission such as a truck with a large total vehicle weight, it is possible to reduce the interstage ratio in deceleration near the direct connection, which is frequently used and difficult to make a small interstage ratio. The degree of freedom of selection of the gear ratio is large, and it is expected that the acceleration performance and fuel efficiency can be improved by selecting the optimum gear ratio according to the driving conditions and running.

In the above-described embodiment, a specific gear train is applied to the first planetary gear group on the upstream side. However, the first planetary gear group 14 is not limited to the above and includes a plurality of speed increases in the forward rotation direction. Any configuration that obtains a gear ratio may be used.
Therefore, it goes without saying that the gear train conventionally proposed or implemented can be applied to the first planetary gear group.

  Further, the second planetary gear group includes the four rotating members as described above, and is a gear train other than the above-described embodiment as long as it has a connection relationship in which the order of arrangement of the speed axes is the same in the common speed diagram. Since the same multi-speed shift is possible, the optimum gear train can be selected based on the general knowledge of those skilled in the art, and the layout or the like can be changed or improved.

  In addition to obtaining a multiple gear ratio with a large gear ratio width, it is possible to reduce the gear ratio at the speed reduction gear that is frequently used and close to the direct gear connection, which is difficult to achieve a small gear ratio. The present invention can be applied to automatic transmissions such as large trucks, and can also be widely applied to transmissions such as industrial vehicles.

It is the skeleton figure which showed the structure of the multi-stage planetary gear train. (Example 1) It is a figure which shows the operation | movement table | surface of Example 1. FIG. It is a figure which shows the common speed diagram (a) thru | or (e) in the 1st planetary gear group of Example 1. FIG. It is a figure which shows the common speed diagram in the 2nd planetary gear group of Example 1. FIG. It is the skeleton figure which showed the structure of the multi-stage planetary gear train. (Example 2) It is a figure which shows the operation | movement table | surface of Example 2. FIG. It is the skeleton figure which showed the structure of the multi-stage planetary gear train. Example 3 It is a figure which shows the operation | movement table | surface of Example 3. FIG. It is a figure which shows the common velocity diagram of Example 3. FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10 Input shaft 12 Output shaft 14 1st gear group 16 1st planetary gear set 18 2nd planetary gear set 20 2nd planetary gear group 22 3rd planetary gear set 24 4th planetary gear set 30 1st sun gear 32 1st ring gear 34 First pinion 38 First carrier 40 Second sun gear 42 Second ring gear 44 Second pinion 48 Second carrier 50 Third sun gear 52 Third ring gear 54 Third pinion 58 Third carrier 60 Fourth sun gear 62 Fourth ring gear 64 Fourth Pinyon 68 Fourth carrier 70 Fifth sun gear 72 Fifth ring gear 74 Fifth pinion 78 Fifth carrier 80 First clutch 82 Second clutch 84 Third clutch 86 First brake 88 Case 90 One-way clutch, First one-way clutch 92 Second Brake 94 3rd brake 96 4th blur Key 98 second one-way clutch 90
100 5th brake 102 4th clutch

Claims (6)

An input shaft;
An output shaft;
An upstream first gear group and a downstream second gear group that are provided between the input shaft and the output shaft and convert the rotational speed of the input shaft to the rotational speed of the output shaft;
The first gear group has an intermediate output member that inputs the output of the first gear group to the second planetary gear group as a rotating member, and in a forward rotation direction between the input shaft and the intermediate output member. It is a configuration to obtain a plurality of gear ratios including speed increase,
The second gear group is composed of two planetary gear sets, and as rotating members, an intermediate input member that can be connected to or connected to the intermediate output member, a low-speed stage fixed member that can be fixed to a case, and the input shaft A high-speed input member that can be connected, and an output member connected to the output shaft;
In the common speed diagram of the second planetary gear group, the speed axes of the intermediate input member, the high speed stage input member, the output member, and the low speed stage fixed member are arranged in this order. Multi-speed planetary gear train.   The first gear group is composed of two or three planetary gear sets, and can reduce at least two stages in the forward rotation direction between the input shaft and the intermediate output member, and at least in the reverse rotation direction. The multi-speed planetary gear train according to claim 1, wherein one-stage reduction is possible. The first gear group includes:
A first planetary gear set having a first sun gear, a first ring gear, a first pinion meshing with the first sun gear and the first ring gear, and a first carrier rotatably supporting the first pinion;
A second planetary gear set having a second sun gear, a second ring gear, a second pinion meshing with the second sun gear and the second ring gear, and a second carrier that rotatably supports the second pinion;
A third planetary gear set having a third sun gear, a third ring gear, a third pinion meshing with the third sun gear and the third ring gear, and a third carrier for rotatably supporting the third pinion. ,
The input shaft can be connected to or connectable to the second sun gear, and can be selectively connected to one of the first carrier, the second ring gear, and the first sun gear connected to each other. Yes,
The first carrier and the second ring gear can be fixed to the case;
The first ring gear and the third carrier are connected to form the intermediate output member;
3. The multi-stage planetary gear train according to claim 1, wherein the first sun gear and the third sun gear can be fixed to the case. 4.   4. The multi-speed planetary gear train according to claim 3, wherein the first sun gear can always be fixed to the case in one rotational direction via a one-way clutch. The first gear group includes:
A first planetary gear set having a first sun gear, a first ring gear, a first pinion meshing with the first sun gear and the first ring gear, and a first carrier rotatably supporting the first pinion;
A second planetary gear set having a second sun gear, a second ring gear, a second pinion meshing with the second sun gear and the second ring gear, and a second carrier rotatably supporting the second pinion. ,
The input shaft is selectively connectable to the second sun gear, the first sun gear, and the first carrier and the second ring gear connected to each other,
The first carrier, the second ring gear, and the first sun gear can be fixed to the case, respectively.
The multi-stage planetary gear train according to claim 1, wherein the first ring gear and the second carrier connected to each other constitute an intermediate output member. The second gear group is
A fourth planetary gear set having a fourth sun gear, a fourth ring gear, a fourth pinion meshing with the fourth sun gear and the fourth ring gear, and a fourth carrier rotatably supporting the fourth pinion;
A fifth planetary gear set having a fifth sun gear, a fifth ring gear, a fifth pinion meshing with the fifth sun gear and the fifth ring gear, and a fifth carrier that rotatably supports the fifth pinion. ,
The fifth ring gear constitutes the intermediate input member;
The fifth carrier and the fourth ring gear connected to each other constitute the high speed stage input member,
The fourth sun gear is connected to or connectable with the fifth sun gear to constitute the low speed stage fixed member;
6. The multi-speed planetary gear train according to claim 1, wherein the fourth carrier constitutes the output member.

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