JP2011117566A - Automatic transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an automatic transmission advantageous in terms of gear efficiency, gear noise, duration reliability and cost while attaining eight forward speeds by three planetary and six friction elements. <P>SOLUTION: The automatic transmission includes a single pinion first planetary gear PG1, a single pinion third planetary gear PG3 and a double pinion second planetary gear PG2. At least eight forward speeds and one reverse speed are attained by a combination of simultaneous engagements of two of six friction elements composed of a first clutch C1 selectively connecting a carrier PC1 with a rotary member M1, a second clutch C2 selectively connecting the carrier PC1 with a ring gear R2, a third clutch C3 selectively connecting a ring gear R1 with the rotary member M1, a fourth clutch C4 selectively connecting the ring gear R1 with a rotary member M2, a first brake B1 for locking the rotation of the rotary member M2, and a second brake B2 for locking the rotation of the ring gear R2. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to an automatic transmission that is applied as a transmission device for a vehicle that has a request for multiple gears and a wide gear ratio.

  Conventionally, as an automatic transmission that achieves a forward 8-speed with three planets and six friction elements, a double pinion type planetary gear and a ravinio type planetary gear unit (one double pinion type planetary gear and one single pinion type planetary gear) ), Four clutches, and two brakes are known (see, for example, Patent Document 1).

JP 2001-182785 A

However, although the conventional automatic transmission achieves a forward 8-speed gear stage with 3 planetary and 6 friction elements, it uses substantially two double pinion type planetary gears. There was a problem of being disadvantageous.
-Gear efficiency and gear noise are poor because the number of gear meshing increases.
・ Since the pinion gear diameter is reduced, durability reliability is reduced.
・ The number of parts increases, which increases costs.

  The present invention has been made by paying attention to the above problems, and an automatic transmission that is advantageous in terms of gear efficiency, gear noise, durability reliability, and cost while achieving eight forward speeds with three planetary six friction elements. The purpose is to provide.

In order to achieve the above object, an automatic transmission according to the present invention includes:
A first planetary gear comprising a first sun gear, a first ring gear, and a first carrier that supports the first sun gear and a first single pinion that meshes with the first ring gear;
A second planetary gear comprising a second sun gear, a second ring gear, and a second carrier that supports the second sun gear and a second double pinion that meshes with the second ring gear;
A third planetary gear comprising a third sun gear, a third ring gear, a third carrier that supports the third sun gear and a third single pinion that meshes with the third ring gear;
6 friction elements,
By appropriately fastening and releasing the six friction elements, it is possible to shift to at least a forward 8-speed gear stage and output torque from the input shaft to the output shaft.
In this automatic transmission,
The input shaft is always connected to the first carrier,
The output shaft is always connected to the third carrier,
The first sun gear is fixed at all times to form a first fixed member,
The second sun gear and the third sun gear are always connected to form a first rotating member,
The second carrier and the third ring gear are always fastened to constitute a second rotating member,
The six friction elements are:
A first friction element that selectively couples between the first carrier and the first rotating member;
A second friction element that selectively couples between the first carrier and the second ring gear;
A third friction element that selectively connects between the first ring gear and the first rotating member;
A fourth friction element that selectively connects between the first ring gear and the second rotating member;
A fifth friction element capable of locking the rotation of the second rotating member;
A sixth friction element capable of locking the rotation of the second ring gear;
Composed of
Of the six friction elements, at least eight forward speeds and one reverse speed are achieved by a combination of two simultaneous fastenings.

Therefore, in the automatic transmission according to the present invention, at least the eight forward speeds and the first reverse speed are achieved by the three planetary and six friction elements. Of these, two single pinion planetary gears and one double pinion planetary gear are used for three planets. For this reason, the number of gear meshes involved in torque transmission is reduced to less than the number of gear meshes when one single pinion planetary gear and two double pinion planetary gears are used, and gear efficiency is improved and gear noise is reduced. To do. And since the gear diameter of a pinion becomes large, durability reliability improves. Furthermore, since the number of parts is reduced, the cost is reduced.
As a result, while achieving eight forward speeds with three planets and six friction elements, it is possible to make an advantage in terms of gear efficiency, gear noise, durability reliability, and cost.

1 is a skeleton diagram illustrating an automatic transmission according to a first embodiment. It is a figure which shows the fastening operation | movement table | surface which achieves 8 forward speeds and 1 reverse speeds by the combination of two simultaneous fastenings among the six friction elements in the automatic transmission of the first embodiment. FIG. 6 is a diagram showing a gear meshing frequency table at each of the forward eighth speed in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram showing a shift operation at a first speed (1st) in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram illustrating a shift operation at a second speed (2nd) in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram illustrating a shift operation at a third speed (3rd) in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram showing a shift operation at a fourth speed (4th) in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram showing a shift operation at a fifth speed (5th) in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram illustrating a shift operation at a sixth speed (6th) in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram showing a shift operation at a seventh speed (7th) in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram showing a shift operation at an eighth speed (8th) in the automatic transmission according to the first embodiment. FIG. 6 is an explanatory diagram illustrating a shift operation at a reverse speed (Rev) in the automatic transmission according to the first embodiment. It is a skeleton figure which shows the automatic transmission of a prior art example. It is a figure which shows the fastening operation | movement table | surface which achieves 8 forward speeds and 2 reverse speeds by the combination of two simultaneous fastenings among six friction elements in the automatic transmission of the conventional example. It is a figure which shows the gear-engagement frequency | count table | surface in each gear stage of 8 forward speeds in the automatic transmission of a prior art example.

  Hereinafter, the best mode for realizing an automatic transmission according to the present invention will be described based on a first embodiment shown in the drawings.

First, the configuration will be described.
FIG. 1 is a skeleton diagram illustrating an automatic transmission according to a first embodiment. Hereinafter, the planetary gear configuration and the friction element configuration of the automatic transmission according to the first embodiment will be described with reference to FIG.

  As shown in FIG. 1, the automatic transmission according to the first embodiment includes a first planetary gear PG1, a second planetary gear PG2, a third planetary gear PG3, an input shaft IN, an output shaft OUT, A first fixed member F1, a first rotating member M1, a second rotating member M2, a first clutch C1 (first friction element), a second clutch C2 (second friction element), Third clutch C3 (third friction element), fourth clutch C4 (fourth friction element), first brake B1 (fifth friction element), and second brake B2 (sixth friction element) And a transmission case TC.

  The first planetary gear PG1 is a single pinion type planetary gear, and includes a first sun gear S1, a first carrier PC1 that supports the first single pinion P1 meshing with the first sun gear S1, and the first And a first ring gear R1 meshing with one single pinion P1.

  The second planetary gear PG2 is a double pinion type planetary gear having second double pinions P2s and P2r, and meshes with the second sun gear S2, the pinion P2s engaged with the second sun gear S2, and the pinion P2s. It consists of a second carrier PC2 that supports the pinion P2r and a second ring gear R2 that meshes with the pinion P2r.

  The third planetary gear PG3 is a single pinion type planetary gear, and includes a third sun gear S3, a third carrier PC3 that supports a third single pinion P3 that meshes with the third sun gear S3, 3 and a third ring gear R3 meshing with the single pinion P3.

  The input shaft IN is a shaft to which rotational drive torque from a drive source (engine or the like) is input via a torque converter or the like, and is always connected to the first carrier PC1.

  The output shaft OUT is a shaft that outputs a rotational drive torque after shifting to the drive wheels via a propeller shaft, a final gear, or the like, and is always connected to the third carrier PC3.

  The first fixed member F1 is a member that always fixes the first sun gear S1 to the transmission case TC.

  The first rotating member M1 is a rotating member that always connects the second sun gear S2 and the third sun gear S3 without interposing a friction element.

  The second rotating member M2 is a rotating member that always fastens the second carrier PC2 and the third ring gear R3 without interposing a friction element.

  The first clutch C1 is a first friction element that selectively connects the first carrier PC1 and the first rotating member M1.

  The second clutch C2 is a second friction element that selectively connects the first carrier PC1 and the second ring gear R2.

  The third clutch C3 is a third friction element that selectively connects the first ring gear R1 and the first rotating member M1.

  The fourth clutch C4 is a fourth friction element that selectively connects the first ring gear R1 and the second rotating member M2.

  The first brake B1 is a fifth friction element that can lock the rotation of the second rotating member M2 with respect to the transmission case TC.

  The second brake B2 is a sixth friction element that can lock the rotation of the second ring gear R2 with respect to the transmission case TC. The second brake B2 is set at an upstream position on the drive source side from the first planetary gear PG1.

  As shown in FIG. 1, the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are connected to the output shaft OUT from the input shaft IN to which a drive source is connected. They are arranged vertically in order.

  FIG. 2 is a diagram illustrating a fastening operation table that achieves eight forward speeds and one reverse speed by combining two of the six friction elements in the automatic transmission according to the first embodiment. FIG. 3 is a diagram showing a gear meshing frequency table at each of the eight forward speeds in the automatic transmission according to the first embodiment. Hereinafter, based on FIG.2 and FIG.3, the transmission structure which establishes each gear stage of the automatic transmission of Example 1 is demonstrated.

  In the automatic transmission of the first embodiment, each of the eight forward speeds and the one reverse speed is set as described below by combining two of the six friction elements C1, C2, C3, C4, B1, and B2 simultaneously. To achieve.

  As shown in FIG. 2, the first speed (1st) is achieved by simultaneously engaging the first clutch C1 and the first brake B1. As shown in FIG. 3, the number of gear meshes at the first speed gear stage is only the third planetary gear PG3 involved in meshing, so the total number of times is 2 times (= 0 times + 0 times + 2 times). It becomes.

  As shown in FIG. 2, the second speed (2nd) is achieved by simultaneous engagement of the third clutch C3 and the first brake B1. As shown in FIG. 3, since the first planetary gear PG1 and the third planetary gear PG3 are involved in meshing, the total number of gear engagements at the second speed gear stage is 4 (= 2). Times + 0 times + 2 times).

  As shown in FIG. 2, the third speed (3rd) is achieved by simultaneously engaging the second clutch C2 and the first brake B1. As shown in FIG. 3, since the second planetary gear PG2 and the third planetary gear PG3 are involved in meshing, the total number of gear engagements at the third speed gear stage is 5 (= 0). Times + 3 times + 2 times).

  As shown in FIG. 2, the fourth speed (4th) is achieved by simultaneous engagement of the second clutch C2 and the third clutch C3. As shown in FIG. 3, the number of gear meshes at the fourth speed gear stage is the total number of times because the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are involved in meshing. Is 7 times (= 2 times + 3 times + 2 times).

  The fifth speed (5th) is achieved by simultaneous engagement of the first clutch C1 and the second clutch C2, as shown in FIG. As shown in FIG. 3, the number of gear meshes at the fifth speed gear stage is such that none of the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 is involved in meshing. The total number of times is zero.

  As shown in FIG. 2, the sixth speed (6th) is achieved by simultaneous engagement of the second clutch C2 and the fourth clutch C4. As shown in FIG. 3, the number of gear meshes at the sixth speed is the total number of times because the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are involved in meshing. Is 7 times (= 2 times + 3 times + 2 times).

  The seventh speed (7th) is achieved by simultaneously engaging the first clutch C1 and the fourth clutch C4 as shown in FIG. As shown in FIG. 3, since the first planetary gear PG1 and the third planetary gear PG3 are involved in the meshing as shown in FIG. Times + 0 times + 2 times).

  As shown in FIG. 2, the eighth speed (8th) is achieved by simultaneous engagement of the third clutch C3 and the fourth clutch C4. As shown in FIG. 3, the number of gear meshes at the eighth speed gear stage is 2 (= 2 times + 0 times + 0 times) since only the first planetary gear PG1 is involved in meshing. It becomes.

  The reverse speed (Rev) is achieved by simultaneous engagement of the first clutch C1 and the second brake B2, as shown in FIG.

Next, the operation will be described.
The operation of the automatic transmission according to the first embodiment will be described by dividing it into “shift operation at each shift stage” and “advantage by comparison with the prior art”.

[Shifting action at each gear stage]
(First gear)
At the first speed (1st), the first clutch C1 and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The input shaft IN, the first carrier PC1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the first clutch C1 and the first rotating member M1. The second carrier PC2 and the third ring gear R3 are fixed to the transmission case TC by the engagement of the first brake B1 and the second rotating member M2.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation to the third sun gear S3 is decelerated and output from the third carrier PC3 in the third planetary gear PG3 fixed to the ring gear. The output rotation speed from the third carrier PC3 (= deceleration rotation speed lower than the input rotation speed) is transmitted as it is to the output shaft OUT, and the first speed gear stage is achieved.

(2nd gear)
At the second speed (2nd), the third clutch C3 and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The first ring gear R1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the third clutch C3 and the first rotating member M1. The second carrier PC2 and the third ring gear R3 are fixed to the transmission case TC by the engagement of the first brake B1 and the second rotating member M2.

  Therefore, when the input rotational speed is input to the first carrier PC1 through the input shaft IN, the input rotation speed is increased and output from the first ring gear R1 in the first planetary gear PG1 fixed with the sun gear. . The rotation of the first ring gear R1 is directly input to the third sun gear S3 via the third clutch C3 and the first rotating member M1. Therefore, in the third planetary gear PG3 fixed to the ring gear, the rotation of the third sun gear S3 is reduced and output from the third carrier PC3. The output rotation speed from the third carrier PC3 (= deceleration rotation speed lower than the input rotation speed and higher than the first speed) is directly transmitted to the output shaft OUT, and the second speed gear stage is achieved.

(3rd speed)
At the third speed (3rd), the second clutch C2 and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By engaging the second clutch C2, the input shaft IN, the first carrier PC1, and the second ring gear R2 are directly connected. The second carrier PC2 and the third ring gear R3 are fixed to the transmission case TC by the engagement of the first brake B1 and the second rotating member M2.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation to the second ring gear R2 is accelerated in the second planetary gear PG2 fixed to the carrier and output from the second sun gear S2. The rotation of the second sun gear S2 is inputted as it is to the third sun gear S3 via the first rotating member M1. Therefore, in the third planetary gear PG3 fixed to the ring gear, the rotation of the third sun gear S3 is reduced and output from the third carrier PC3. The output rotation speed (= deceleration rotation speed lower than the input rotation speed and higher than the second speed) from the third carrier PC3 is directly transmitted to the output shaft OUT, and the third speed gear stage is achieved.

(4th gear)
At the fourth speed (4th), the second clutch C2 and the third clutch C3 are simultaneously engaged as shown by hatching in FIG.

  By engaging the second clutch C2, the input shaft IN, the first carrier PC1, and the second ring gear R2 are directly connected. The first ring gear R1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the third clutch C3 and the first rotating member M1.

  Therefore, when the input rotational speed is input to the first carrier PC1 through the input shaft IN, the input rotation speed is increased and output from the first ring gear R1 in the first planetary gear PG1 fixed with the sun gear. . The rotation of the first ring gear R1 is directly input to the second sun gear S2 and the third sun gear S3 via the third clutch C3 and the first rotating member M1. For this reason, in the second planetary gear PG2 having two inputs and one output, the rotation speed of the second sun gear S2 and the rotation speed of the second ring gear R2 (= input rotation speed) are defined, whereby the second carrier The rotation speed of PC2 is determined. The rotation of the second carrier PC2 is directly input to the third ring gear R3 via the second rotating member M2. Therefore, in the third planetary gear PG3 having two inputs and one output, the rotational speed of the third sun gear S3 (= the rotational speed of the first ring gear R1) and the rotational speed of the third ring gear R3 are defined. The number of rotations of the third carrier PC3 is determined. The output rotation speed from the third carrier PC3 (deceleration rotation speed lower than the input rotation speed and higher than the third speed) is directly transmitted to the output shaft OUT, and the fourth speed gear stage is achieved.

(5th speed)
At the fifth speed (5th), the first clutch C1 and the second clutch C2 are simultaneously engaged as shown by hatching in FIG.

  By the simultaneous engagement of the first clutch C1 and the second clutch C2 and the first and second rotating members M1 and M2, the two planetary gears PG2 directly connect the two rotating elements S2 and R2 to the second planetary gear. Three rotation elements S2, PC2, and R2 of PG2 are rotated together, and two rotation elements S3 and R3 are directly connected to each other in the third planetary gear PG3, so that three rotations of the third planetary gear PG3 are performed. The elements S3, PC3, and R3 are rotated together, and the input shaft IN, the first carrier PC1, the second planetary gear PG2, and the third planetary gear PG3 are directly connected.

  Therefore, when the input shaft IN rotates at the input rotation speed, the third planetary gear PG3 rotates integrally at the input rotation speed. For this reason, the output rotational speed from the third carrier PC3 (the same rotational speed as the input rotational speed from the input shaft IN) is transmitted to the output shaft OUT as it is, and the fifth speed gear stage (directly coupled speed change) with a gear ratio 1 Stage) is achieved.

(Sixth speed gear)
At the sixth speed (6th), the second clutch C2 and the fourth clutch C4 are simultaneously engaged as shown by hatching in FIG.

  By engaging the second clutch C2, the input shaft IN, the first carrier PC1, and the second ring gear R2 are directly connected. The first ring gear R1, the second carrier PC2, and the third ring gear R3 are directly connected by the engagement of the fourth clutch C4 and the second rotating member M2.

  Accordingly, when the input rotational speed is input to the first carrier PC1 through the input shaft IN, the input rotational speed is increased and output from the first ring gear R1 in the first planetary gear PG1 fixed to the sun gear. . The rotation of the first ring gear R1 is directly input to the second carrier PC2 and the third ring gear R3 via the fourth clutch C4 and the second rotating member M2. For this reason, in the second planetary gear PG2 having two inputs and one output, the rotation speed of the second carrier PC2 and the rotation speed of the second ring gear R2 (= input rotation speed) are defined, whereby the second sun gear. The rotation speed of S2 is determined. The rotation of the second sun gear S2 is directly input to the third sun gear S3 via the first rotating member M1. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotational speed of the third sun gear S3 and the rotational speed of the third ring gear R3 (= the rotational speed of the first ring gear R1) are defined. The rotational speed of the third carrier PC3 is determined. The output rotation speed from the third carrier PC3 (acceleration speed higher than the input rotation speed) is transmitted as it is to the output shaft OUT, and the sixth speed is achieved.

(7th speed)
At the seventh speed (7th), the first clutch C1 and the fourth clutch C4 are simultaneously engaged as shown by hatching in FIG.

  The input shaft IN, the first carrier PC1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the first clutch C1 and the first rotating member M1. The first ring gear R1, the second carrier PC2, and the third ring gear R3 are directly connected by the engagement of the fourth clutch C4 and the second rotating member M2.

  Therefore, when the input shaft IN is rotated at the input rotational speed, the input rotation to the first carrier PC1 is increased in the first planetary gear PG1 fixed to the sun gear, and output from the first ring gear R1. The rotation of the first ring gear R1 is directly input to the third ring gear R3 via the fourth clutch C4 and the second rotating member M2. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotation speed of the third sun gear S3 (= input rotation speed) and the rotation speed of the third ring gear R3 are defined, so that the third carrier The rotation speed of PC3 is determined. The output rotational speed (the input rotational speed and the speed increasing rotational speed higher than the sixth speed) from the third carrier PC3 is directly transmitted to the output shaft OUT, and the seventh speed gear stage is achieved.

(8th speed)
At the eighth speed (8th), the third clutch C3 and the fourth clutch C4 are simultaneously engaged, as shown by hatching in FIG.

  By the simultaneous engagement of the third clutch C3 and the fourth clutch C4 and the first and second rotating members M1 and M2, the two planetary gears PG2 directly connect the two rotating elements S2 and PC2 to the second planetary gear. Three rotation elements S2, PC2, and R2 of PG2 are brought into a state of rotating integrally, and two rotation elements S3 and R3 are directly connected in the third planetary gear PG3, and three rotations of the third planetary gear PG3 are performed. The elements S3, PC3, and R3 are rotated together, and the first ring gear R1, the second planetary gear PG2, and the third planetary gear PG3 are directly connected.

  Therefore, when the input rotational speed is input to the first carrier PC1 through the input shaft IN, the input rotation speed is increased and output from the first ring gear R1 in the first planetary gear PG1 fixed with the sun gear. . The rotation of the first ring gear R1 is directly input to the third planetary gear PG3 via the third clutch C3, the fourth clutch C4, and the first and second rotating members M1 and M2, and the third planetary gear. PG3 rotates integrally with the rotation of the first ring gear R1. For this reason, the output rotational speed (the input rotational speed and the speed increasing rotational speed higher than the seventh speed) from the third carrier PC3 is directly transmitted to the output shaft OUT, and the eighth speed gear stage is achieved.

(Reverse speed shift stage)
At the reverse speed (Rev), the first clutch C1 and the second brake B2 are simultaneously engaged as shown by hatching in FIG.

  The input shaft IN, the first carrier PC1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the first clutch C1 and the first rotating member M1. As the second brake B2 is engaged, the second ring gear R2 is fixed to the transmission case TC.

  Therefore, when the input shaft IN rotates at the input rotation speed, the second planetary gear PG2 fixed to the ring gear outputs a rotation in the opposite direction to the input rotation to the second sun gear S2 from the second carrier PC2. The rotation of the second carrier PC2 is directly input to the third ring gear R3 via the second rotating member M2. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotation speed of the third sun gear S3 (= input rotation speed) and the rotation speed of the third ring gear R3 are defined, so that the third carrier The rotation speed of PC3 is determined. The output rotational speed from the third carrier PC3 (a decelerating rotational speed opposite to the input rotational speed and lower than the input rotational speed) is transmitted as it is to the output shaft OUT, and a reverse speed gear stage is achieved.

[Advantages by comparison with conventional technology]
FIG. 13 is a skeleton diagram showing a conventional automatic transmission. FIG. 14 is a diagram showing a fastening operation table for achieving 8 forward speeds and 2 reverse speeds by combining two of the six friction elements in the conventional automatic transmission. Hereinafter, the advantages of the automatic transmission according to the first embodiment in comparison with the prior art will be described with reference to FIGS. 13 and 14.

First, when comparing the automatic transmission of the first embodiment (FIGS. 1 and 2) and the automatic transmission of the conventional example (FIGS. 13 and 14), it can be said that the performance is equivalent with respect to the points listed below. .
(Basic configuration and speed change performance)
Both the automatic transmission of the conventional example and the automatic transmission of the first embodiment achieve eight forward speeds and one reverse speed with three planets and six friction elements.
(Transmission control performance)
In the automatic transmission of the conventional example and the automatic transmission of the first embodiment, the shift to the adjacent shift stage and the shift to the one-step jump shift stage are called release of one friction element and engagement of one friction element. This is achieved by a single overhanging shift.

  However, “(a) 3 planetary gears” “(b) gear ratio width” “(c) reverse power performance” “(d) unit layout” “(e) automatic transmission unit shape” “(f The automatic transmission according to the first embodiment has an advantage over the automatic transmission according to the conventional example.

(a) When selecting a planetary gear to be used for a three planetary gear automatic transmission, there are single pinion type planetary gears and double pinion type planetary gears as options, but from the viewpoint of gear transmission efficiency, etc., double pinion type planetary gears However, it is considered preferable to select a single pinion type planetary gear.

  As shown in FIG. 13, the automatic transmission of the conventional example uses a double pinion type planetary gear and a Ravinio type planetary gear unit (one double pinion type planetary gear and one single pinion type planetary gear). That is, since two double pinion type planetary gears are used substantially, the number of gear meshing increases, the transmission efficiency and gear noise of the gear are poor, the pinion gear diameter is reduced, and the durability reliability is reduced. There is a problem that the cost increases because the number of parts increases.

  On the other hand, in the case of the automatic transmission according to the first embodiment, the second planetary gear PG2 using a double pinion, the first planetary gear PG1 and the third planetary gear PG3 using a single pinion are used. For this reason, compared with the conventional example using two double pinion type planetary gears, the number of use of double pinion type planetary gears decreases. For this reason, compared with the conventional example which uses two double pinion type planetary gears, it becomes advantageous in the following items.

In the case of the automatic transmission according to the first embodiment, the number of gear meshes is reduced as compared with the conventional example, the transmission efficiency of the gear is improved, and the gear noise is reduced.
That is, one set of double pinion type planetary gears has a meshing number of 3, whereas one set of single pinion type planetary gears has a meshing number of 2 because there is no meshing between the pinions. Therefore, in the case of Example 1, the average meshing number is 3.88 as shown in FIG. On the other hand, in the case of the conventional example using two pairs of double pinion type planetary gears, the average meshing number is 4.75 as shown in FIG. As a result, the average meshing number is reduced by 0.87, the transmission efficiency of the gear is improved, and the gear noise is reduced.

In the case of the automatic transmission of the first embodiment, the pinion gear diameter is increased, so that the durability reliability is improved.
That is, in the case of a single pinion, a plurality of pinions whose gear diameter is the distance between both gears are arranged between the sun gear and the ring gear. On the other hand, in the case of a double pinion, it is necessary to make the diameter smaller than the distance between the two gears. Thus, in the case of a single pinion, since the gear diameter of the pinion is larger than that of the double pinion, the rigidity and tooth surface strength of the pinion can be increased, and the durability reliability is improved.

In the case of the automatic transmission of the first embodiment, the number of parts is reduced, which is advantageous in terms of cost.
For example, in the case of a double pinion type planetary gear, when four sets of double pinions are arranged around the sun gear, the number of pinions is eight. On the other hand, in the case of a single pinion type planetary gear, it is sufficient to arrange four pinions around the sun gear, and the number of parts is reduced by four. As a result, cost reduction can be achieved.

(b) Gear ratio width The range of change in the gear ratio of the automatic transmission is represented by the ratio coverage (= minimum speed gear ratio / maximum speed gear ratio: hereinafter referred to as “RC”). The larger the RC value, the wider the gear ratio change range, and the higher the gear ratio setting freedom, the better.

  In the case of the conventional automatic transmission, RC = 6.397 (= 4.267 / 0.667), as shown in FIG. On the other hand, in the automatic transmission of the first embodiment, as shown in FIG. 2, the tooth number ratio of the first planetary gear PG1 is ρ1 = 0.630, and the tooth number ratio of the second planetary gear PG2 is ρ2 = -0.390. When the gear ratio of the third planetary gear PG3 is ρ3 = 0.330, RC = 6.574 (= 4.030 / 0.613) is obtained while maintaining an appropriate gear ratio at the adjacent gears.

  In other words, while maintaining an appropriate gear ratio, the RC value can be made larger than the conventional example, ensuring both the starting performance at the lowest gear ratio and the high speed fuel efficiency at the highest gear ratio. can do. Here, “appropriate gear ratio” means that the gear ratio at each gear stage is plotted, and when the plotted points are connected by a line, the characteristic is smooth from the low gear side toward the high gear side. It means that a characteristic line can be drawn that changes in a flat state after being lowered by a gradient.

  The rotational speed actually transmitted to the drive wheels is adjusted by the final gear ratio of the final reduction gear provided at the downstream position of the automatic transmission. Therefore, the higher the RC value, the higher the degree of freedom of adjustment by the final gear ratio. For example, by adjusting to the low side, it is advantageous to cope with an automatic transmission of a hybrid vehicle having no torque converter. Become. In addition, it will be advantageous to handle gasoline engines and diesel engines with different optimal fuel efficiency and maximum torque ranges. That is, in the case of an engine vehicle, both the starting driving force and the fuel efficiency (reduction in engine speed) can be achieved.

(c) Reverse power performance The first gear ratio and reverse gear ratio are values that determine start acceleration and climbing performance. For example, if the ratio between the first gear ratio and the reverse gear ratio is not in the vicinity of 1, the vehicle moves forward and backward. A driving force difference occurs at the time of switching. Also, if the reverse gear ratio is lower than the first gear ratio, the driving force at the time of backward start is lower than the driving force at the time of forward start, and the reverse startability is inferior.

  In the case of the conventional automatic transmission, as shown in FIG. 14, Rev1 / 1st = 0.750 and Rev2 / 1st = 0.469. Therefore, in the case of Rev1 / 1st, that is, if the reverse 1st speed (Rev1) is selected, the level that can prevent the driving force from being insufficient during reverse can be maintained, but if the 2nd reverse speed (Rev2) is selected, the 1st gear The ratio between the ratio and the reverse gear ratio is a value significantly lower than 1, and a driving force difference may occur when switching between forward and reverse travel, and reverse startability may be deteriorated.

  On the other hand, in the automatic transmission of the first embodiment, as shown in FIG. 2, Rev / 1st = 1.067, and the ratio of the first gear ratio and the reverse gear ratio is in the vicinity of 1. For this reason, a driving force difference does not occur at the time of forward / reverse switching, and the backward startability is not deteriorated. That is, the vehicle can be operated without impairing the start acceleration performance and the climbing performance.

(d) Unit layout In the automatic transmission of the first embodiment, the gear ratios ρ1, ρ2, ρ3 of the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are within the range of 0.3 to 0.65. Therefore, it is possible to suppress an increase in the size of each planetary gear PG1 to PG3 while securing the gear tooth surface strength, gear diameter, and number of teeth of each planetary gear PG1 to PG3 in an appropriate state. As a result, expansion of the unit layout can be prevented.

  Further, in the automatic transmission according to the first embodiment, since the maximum torque sharing ratio in each friction element (first clutch C1 to first brake B1) is 1.5 or less, each friction element (first clutch C1 to first brake B1). ) Can be suppressed, and the unit layout can be prevented from expanding.

  By preventing the expansion of the unit layout, the transmission case TC can be made compact, which greatly contributes to miniaturization and weight reduction of the automatic transmission and cost reduction.

(e) Automatic transmission unit shape Among friction elements by clutch elements and brake elements, the brake element is arranged between the rotating element and the transmission case, and if the torque sharing ratio of the brake element is high, the number of plates increases. It is necessary to cope with this by increasing the diameter of the transmission case.

  In the case of the conventional automatic transmission, the second brake B2 is the brake element having the highest torque sharing ratio, and as shown in FIG. 13, the second brake B2 is located between the double pinion planetary gear and the ravinio type planetary gear unit. In order to avoid interference with the vehicle body floor, it is necessary to form a floor tunnel that protrudes greatly into the passenger compartment.

  On the other hand, in the case of the automatic transmission of the first embodiment, the second brake B2 is the brake element having the highest torque sharing ratio, but as shown in FIG. Is disposed at the front side of the first planetary gear PG1, which is irrelevant, that is, at the upstream position on the drive source side of the first planetary gear PG1, so that the body diameter of the case region overlapping the vehicle body floor can be reduced. it can. For this reason, if the body diameter of only the front side portion of the transmission case TC arranged in the power unit room (for example, the engine room) is made thicker, it can be set to a shape setting that makes the body diameter thereafter narrower. It is possible to prevent interference with the vehicle body floor simply by forming a slightly protruding floor tunnel.

(f) Shift Frequency In the conventional automatic transmission, as shown in FIG. 14, the sixth speed is set as the direct coupling stage, and the first to fifth speeds are set as the underdrive shift stage. Therefore, the speed change interval on the underdrive side becomes small, for example, a busy shift with a high shift frequency in an urban area where the stop and the start are repeated, and in the case of an engine vehicle, the engine speed on the underdrive side Since the number rises quickly, the feeling is worsened by the busy shift.

  On the other hand, in the automatic transmission according to the first embodiment, as shown in FIG. 2, the fifth speed is set as a direct coupling stage, and the first to fourth speeds are set as underdrive shift stages. Therefore, since the shift interval on the underdrive side is wider than that in the conventional example, for example, in a city area where the stop and the start are repeated, a busy shift can be suppressed, and deterioration in feeling can be prevented.

Next, the effect will be described.
In the automatic transmission of the first embodiment, the effects listed below can be obtained.

(1) A first sun gear S1, a first ring gear R1, and a first carrier PC1 supporting a first single pinion P1 meshing with the first sun gear S1 and the first ring gear R1. A second planetary gear PG1, a second sun gear S2, a second ring gear R2, and a second double pinion P2s, P2r that meshes with the second sun gear S2 and the second ring gear R2. Supports a second planetary gear PG2 comprising a carrier PC2, a third sun gear S3, a third ring gear R3, and a third single pinion P3 meshing with the third sun gear S3 and the third ring gear R3. A third planetary gear PG3 comprising a third carrier PC3 and six friction elements, and by appropriately fastening and releasing the six friction elements, the speed is shifted to at least a forward 8-speed gear stage and input. Output torque from shaft IN to output shaft OUT In the automatic transmission, the input shaft IN is always connected to the first carrier PC1, the output shaft OUT is always connected to the third carrier PC3, and the first sun gear. S1 is always fixed to form a first fixed member F1, and the second sun gear S2 and the third sun gear S3 are always connected to form a first rotating member M1, The second carrier PC2 and the third ring gear R3 are always fastened to constitute a second rotating member M2, and the six friction elements are the first carrier PC1 and the first rotation. A first friction element (first clutch C1) that selectively connects the members M1 and a second friction element (selective connection between the first carrier PC1 and the second ring gear R2) The second clutch C2), and the selection between the first ring gear R1 and the first rotating member M1 A third friction element (third clutch C3) coupled to the first ring gear, and a fourth friction element (fourth clutch C4) selectively coupled between the first ring gear R1 and the second rotating member M2. , A fifth friction element (first brake B1) that can lock the rotation of the second rotating member M2, and a sixth friction element (second brake) that can lock the rotation of the second ring gear R2. B2), and at least 8 forward speeds and 1 reverse speed are achieved by combining two of the six friction elements at the same time.
Therefore, while achieving 8 forward speeds with 3 planetary 6 friction elements, the number of double pinion planetary gears can be reduced, which is advantageous in terms of gear efficiency, gear noise, durability reliability, and cost. Since the number of friction elements is increased without increasing the number of friction elements, it is possible to improve fuel consumption performance and transmission performance without expanding the unit layout and increasing costs.

(2) Of the six friction elements, the first friction element (first clutch C1) and the fifth friction element (first brake B1) are at least in the eighth forward speed by the combination of two simultaneous engagements. The first speed achieved by simultaneous engagement of the second friction element, the second speed achieved by simultaneous engagement of the third friction element (third clutch C3) and the fifth friction element (first brake B1), and the second speed. The third speed achieved by simultaneous engagement of the second friction element (second clutch C2) and the fifth friction element (first brake B1), the second friction element (second clutch C2) and the third The fourth speed achieved by simultaneous engagement of the friction element (third clutch C3) and the simultaneous engagement of the first friction element (first clutch C1) and the second friction element (second clutch C2). Fifth speed, the second friction element (second clutch C2) and the fourth speed The sixth speed achieved by simultaneous engagement of the friction element (fourth clutch C4), and the simultaneous engagement of the first friction element (first clutch C1) and the fourth friction element (fourth clutch C4). The seventh speed and the eighth speed achieved by simultaneous engagement of the third friction element (third clutch C3) and the fourth friction element (fourth clutch C4) are configured.
For this reason, the shift to the adjacent stage and the one-step jumping shift stage is achieved by single replacement by fastening one friction element and releasing one friction element, which is advantageous because the shift control is simplified. In addition, the RC value can be set so as to reach a required value for achieving both the start performance at the lowest gear ratio and the high speed fuel consumption at the highest gear ratio while maintaining an appropriate gear ratio. In addition, by setting the direct coupling stage to the fifth speed, it is possible to widen the shift interval on the low speed stage side and to prevent busy shift.

(3) Of the six friction elements, the first reverse speed achieved by the combination of two simultaneous engagements is the first friction element (first clutch C1) and the sixth friction element (second brake B2). The configuration is achieved by simultaneous fastening.
Therefore, the reverse gear ratio evaluation value (= reverse gear ratio / 1st gear ratio) can be set to a value near 1 even if a gear ratio that achieves an appropriate RC value and interstage ratio is selected. As a result, it is possible to prevent the occurrence of a driving force difference when switching between forward and backward travel, and it is possible to ensure reverse start acceleration and climbing performance.

(4) The first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are vertically arranged in order from the input shaft IN to which the drive source is connected to the output shaft OUT. The sixth friction element (second brake B2) is set at the upstream position on the drive source side from the first planetary gear PG1.
For this reason, by increasing the body diameter of only the front part of the transmission case TC and making the body diameter of the subsequent part thinner, it is possible to enter the floor tunnel cabin while preventing interference with the vehicle body floor. Can be kept small.

  As mentioned above, although the automatic transmission of this invention has been demonstrated based on Example 1, it is not restricted to this Example 1 about a concrete structure, The summary of the invention which concerns on each claim of a claim As long as they do not deviate, design changes and additions are permitted.

  In the first embodiment, the tooth number ratio ρ1 of the first planetary gear PG1, the tooth number ratio ρ2 of the second planetary gear PG2, and the tooth number ratio ρ3 of the third planetary gear PG3 are set to preferable values. An example of setting is shown. However, the gear ratios ρ1, ρ2, and ρ3 of the planetary gears PG1, PG2, and PG3 are values within a range in which the gear ratio can be set, and a gear ratio with a high RC value and an appropriate interstage ratio are obtained. The specific values are not limited to the values in the first embodiment as long as they are set as described above.

  In the first embodiment, an example of an automatic transmission that is applied to an FR engine vehicle in which the input / output shaft is coaxially arranged is shown. However, the automatic transmission is not limited to an FR engine vehicle, but an FF engine vehicle, a hybrid vehicle, an electric vehicle, and a fuel cell vehicle. The present invention can also be applied as an automatic transmission for various vehicles.

PG1 first planetary gear
PG2 Second planetary gear
PG3 3rd planetary gear
IN input shaft
OUT output shaft
F1 First fixed member
M1 First rotating member
M2 Second rotating member
C1 First clutch (first friction element)
C2 Second clutch (second friction element)
C3 3rd clutch (3rd friction element)
C4 4th clutch (4th friction element)
B1 First brake (fifth friction element)
B2 Second brake (sixth friction element)
TC transmission case

Claims (4)

A first planetary gear comprising a first sun gear, a first ring gear, and a first carrier that supports the first sun gear and a first single pinion that meshes with the first ring gear;
A second planetary gear comprising a second sun gear, a second ring gear, and a second carrier that supports the second sun gear and a second double pinion that meshes with the second ring gear;
A third planetary gear comprising a third sun gear, a third ring gear, a third carrier that supports the third sun gear and a third single pinion that meshes with the third ring gear;
6 friction elements,
In an automatic transmission capable of shifting to at least a forward 8-speed gear stage by appropriately fastening and releasing the six friction elements and outputting torque from the input shaft to the output shaft,
The input shaft is always connected to the first carrier,
The output shaft is always connected to the third carrier,
The first sun gear is fixed at all times to form a first fixed member,
The second sun gear and the third sun gear are always connected to form a first rotating member,
The second carrier and the third ring gear are always fastened to constitute a second rotating member,
The six friction elements are:
A first friction element that selectively couples between the first carrier and the first rotating member;
A second friction element that selectively couples between the first carrier and the second ring gear;
A third friction element that selectively connects between the first ring gear and the first rotating member;
A fourth friction element that selectively connects between the first ring gear and the second rotating member;
A fifth friction element capable of locking the rotation of the second rotating member;
A sixth friction element capable of locking the rotation of the second ring gear;
Composed of
An automatic transmission characterized in that at least 8 forward speeds and 1 reverse speed are achieved by a combination of two simultaneous engagements among the six friction elements. The automatic transmission according to claim 1, wherein
Of the six friction elements, at least eight forward speeds are achieved by a combination of two simultaneous fastenings.
A first speed achieved by simultaneous engagement of the first friction element and the fifth friction element;
A second speed achieved by simultaneous engagement of the third friction element and the fifth friction element;
A third speed achieved by simultaneous engagement of the second friction element and the fifth friction element;
A fourth speed achieved by simultaneous engagement of the second friction element and the third friction element;
A fifth speed achieved by simultaneous engagement of the first friction element and the second friction element;
A sixth speed achieved by simultaneous engagement of the second friction element and the fourth friction element;
A seventh speed achieved by simultaneous engagement of the first friction element and the fourth friction element;
An eighth speed achieved by simultaneous engagement of the third friction element and the fourth friction element;
An automatic transmission characterized by comprising: In the automatic transmission according to claim 1 or 2,
An automatic transmission characterized in that the first reverse speed achieved by a combination of two simultaneous engagements among the six friction elements is achieved by simultaneous engagement of the first friction element and the sixth friction element. In the automatic transmission according to any one of claims 1 to 3,
The first planetary gear, the second planetary gear, and the third planetary gear are vertically arranged in order from the input shaft to which the drive source is connected toward the output shaft,
An automatic transmission characterized in that the sixth friction element is set at an upstream position on the drive source side from the first planetary gear.

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