JP5021009B2 - Automatic transmission - Google Patents

Description

  The present invention relates to an automatic transmission that is applied as a transmission device for a vehicle that has a request for a multi-stage shift speed and a request for a wide gear ratio width.

  Conventionally, as an automatic transmission that achieves a forward eight-speed shift stage with three planets and six friction elements, a double pinion type planetary gear, a Rabinio type planetary gear unit (one double pinion planet and one single pinion planet), One having four clutches and two brakes is known (for example, see 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.

  Further, in order to achieve each of the eight forward speeds, two friction elements are fastened. Therefore, there are four idling friction elements at each gear stage, resulting in friction loss in the idling friction elements. However, there is a problem that the transmission efficiency of drive energy is deteriorated.

  In particular, in the case of multi-plate clutches and multi-plate brakes that are frequently used as friction elements, oil that is sprayed for cooling or lubrication is interposed between the relatively rotating plates in the idling state due to element release, and drag resistance (oil The generation of friction loss due to the shear resistance) cannot be avoided. Moreover, the friction loss increases as the number of plates increases and the relative rotational speed between the plates increases.

  The present invention has been made by paying attention to the above-mentioned problems, and is advantageous in terms of gear efficiency, gear noise, durability reliability, and cost while achieving eight forward speeds with three planetary six friction elements and reduces friction loss. An object of the present invention is to provide an automatic transmission capable of improving the transmission efficiency of drive energy by keeping it small.

In order to achieve the above object, an automatic transmission according to the present invention supports a first sun gear, a first ring gear, and a first double pinion that meshes with the first sun gear and the first ring gear. A first planetary gear comprising:
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 single pinion meshing 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 second ring gear,
The second carrier and the third carrier are always connected to form a first rotating member,
The second sun gear and the third sun gear are always connected to form a second rotating member,
The six friction elements are:
A first friction element that selectively connects between the first sun gear and the second rotating member;
A second friction element that selectively connects between the first ring gear and the second rotating member;
A third friction element that selectively connects between the first ring gear and the third ring gear;
A fourth friction element that selectively connects between the first carrier and the first rotating member;
A fifth friction element capable of locking the rotation of the first sun gear;
A sixth friction element capable of locking the rotation of the first rotating member;
Composed of
Of the six friction elements, at least eight forward speeds and one reverse speed are achieved by a combination of three 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, improving gear efficiency and reducing gear noise. 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.
Further, for the six friction elements, each gear stage is achieved by a combination of three simultaneous engagements. For this reason, there are three friction elements that idle in each shift stage, and friction loss at the friction element that idles can be reduced compared to the case where each shift stage is achieved by a combination of two simultaneous engagements. Therefore, for example, when applied to an engine vehicle, drive energy transmission efficiency is improved such that fuel efficiency is improved.
As a result, while achieving 8 forward speeds with 3 planetary 6 friction elements, it is advantageous in terms of gear efficiency, gear noise, durability reliability, and cost, and also improves drive energy transmission efficiency by minimizing friction loss. Can be achieved.

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 speed | velocity | rate by the combination of three simultaneous fastening among the six friction elements in the automatic transmission of Example 1. FIG. 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. It is a skeleton figure which shows the automatic transmission of Example 2. FIG. 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 three simultaneous fastening among the six friction elements in the automatic transmission of the second embodiment. FIG. 10 is a diagram showing a gear meshing frequency table at each of the forward eighth speed in the automatic transmission according to the second embodiment. FIG. 10 is an explanatory diagram showing a shift operation at a first speed (1st) in the automatic transmission according to the second embodiment. FIG. 7 is an explanatory diagram showing a shift operation at a second speed (2nd) in the automatic transmission according to the second embodiment. FIG. 10 is an explanatory diagram showing a shift operation at a third speed (3rd) in the automatic transmission according to the second embodiment. FIG. 12 is an explanatory diagram showing a shift operation at a fourth speed (4th) in the automatic transmission according to the second embodiment. FIG. 12 is an explanatory diagram showing a shift operation at a fifth speed (5th) in the automatic transmission according to the second embodiment. FIG. 12 is an explanatory diagram showing a shift operation at a sixth speed (6th) in the automatic transmission according to the second embodiment. FIG. 12 is an explanatory diagram showing a shift operation at a seventh speed (7th) in the automatic transmission according to the second embodiment. FIG. 12 is an explanatory diagram showing a shift operation at an eighth speed (8th) in the automatic transmission according to the second embodiment. FIG. 12 is an explanatory diagram showing a shift operation at a reverse speed (Rev) in the automatic transmission according to the second embodiment.

  Hereinafter, the best mode for realizing the automatic transmission of the present invention will be described based on Example 1 and Example 2 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, The first rotary member M1, the second rotary member M2, the first clutch C1 (first friction element), the second clutch C2 (second friction element), and the third clutch C3 (third Friction element), fourth clutch C4 (fourth friction element), first brake B1 (fifth friction element), second brake B2 (sixth friction element), and transmission case TC. I have.

  The first planetary gear PG1 is a double pinion type planetary gear, and a first sun gear S1, a first ring gear R1, and a first double pinion P1s meshing with the first sun gear S1 and the first ring gear R1. , P1r for supporting the first carrier PC1.

  The second planetary gear PG2 is a single pinion type planetary gear, and a second single gear P2 meshing with the second sun gear S2, the second ring gear R2, the second sun gear S2 and the second ring gear R2. And a second carrier PC2 that supports the second carrier PC2.

  The third planetary gear PG3 is a single pinion type planetary gear, and 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. And a third carrier PC3 for supporting

  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 for outputting the rotational drive torque after shifting to the drive wheels via a propeller shaft, final gear, or the like, and is always connected to the second ring gear R2.

  The first rotating member M1 is a rotating member that always connects the second carrier PC2 and the third carrier PC3 without interposing a friction element.

  The second rotating member M2 is a rotating member that always connects the second sun gear S2 and the third sun gear S3 without interposing a friction element.

  The first clutch C1 is a first friction element that selectively connects the first sun gear S1 and the second rotating member M2.

  The second clutch C2 is a second friction element that selectively connects the first ring gear R1 and the second rotating member M2.

  The third clutch C3 is a third friction element that selectively connects the first ring gear R1 and the third ring gear R3.

  The fourth clutch C4 is a fourth friction element that selectively connects the first carrier PC1 and the first rotating member M1.

  The first brake B1 is a fifth friction element that can lock the rotation of the first sun gear S1 with respect to the transmission case TC.

  The second brake B2 is a sixth friction element that can lock the rotation of the first rotating member M1 with respect to the transmission case TC.

  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 in order.

  FIG. 2 is a diagram showing a fastening operation table that achieves eight forward speeds and one reverse speed by combining three of the six friction elements in the automatic transmission of 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.

  The automatic transmission according to the first embodiment has eight forward speeds and one reverse speed as described below by combining three 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, the third clutch C3, and the second brake B2. As shown in FIG. 3, since the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are involved in the meshing, the number of gear meshes at the first speed gear stage is the total number of gear meshes. Is 7 times (= 3 times + 2 times + 2 times).

  The second speed (2nd) is achieved by simultaneous engagement of the third clutch C3, the first brake B1, and the second brake B2, as shown in FIG. As shown in FIG. 3, the number of gear meshes at the second 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 (= 3 times + 2 times + 2 times).

  As shown in FIG. 2, the third speed (3rd) is established by simultaneously engaging the first clutch C1, the third clutch C3, and the first brake B1. As shown in FIG. 3, the number of gear meshes at the third 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 (= 3 times + 2 times + 2 times).

  As shown in FIG. 2, the fourth speed (4th) is achieved by simultaneously engaging the second clutch C2, the third clutch C3, and the first brake B1. As shown in FIG. 3, the number of gear meshes at the fourth speed gear stage is 3 (= 3 times + 0 times + 0 times) because only the first planetary gear PG1 is involved in meshing. It becomes.

  As shown in FIG. 2, the fifth speed (5th) is achieved by simultaneously engaging the third clutch C3, the fourth clutch C4, and the first brake B1. As shown in FIG. 3, the number of gear meshes at the fifth 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 (= 3 times + 2 times + 2 times).

  As shown in FIG. 2, the sixth speed (6th) is achieved by simultaneously engaging the second clutch C2, the third clutch C3, and the fourth clutch C4. As shown in FIG. 3, the number of gear meshes at the sixth 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 seventh speed (7th) is achieved by simultaneously engaging the second clutch C2, the fourth clutch C4, and the first brake B1. As shown in FIG. 3, since the first planetary gear PG1 and the second planetary gear PG2 are involved in meshing, the total number of gear engagements at the seventh speed gear stage is 5 (= 3). Times +2 times +0 times).

  As shown in FIG. 2, the eighth speed (8th) is achieved by simultaneously engaging the first clutch C1, the fourth clutch C4, and the first brake B1. As shown in FIG. 3, the number of gear meshes at the eighth speed gear stage is only 2nd planetary gear PG2 is involved in meshing, so the total number of gear meshes is 2 times (= 0 times + 2 times + 0 times). It becomes.

  As shown in FIG. 2, the reverse speed (Rev) is achieved by simultaneously engaging the second clutch C2, the first brake B1, and the second brake B2.

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, the third clutch C3, and the second brake B2 are simultaneously engaged as shown by hatching in FIG.

  The first sun gear S1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the first clutch C1 and the second rotating member M2. By engaging the third clutch C3, the first ring gear R1 and the third ring gear R3 are directly connected. The second carrier PC2 and the third carrier PC3 are fixed to the transmission case TC by the engagement of the second brake B2 and the first rotating member M1.

  Therefore, when the input rotational speed is input to the first carrier PC1 after passing through the input shaft IN, the rotation of the first sun gear S1 and the first ring gear R1 of the first planetary gear PG1 The third planetary gear PG3 rotates while being restrained by the rotation of the third sun gear S3 and the third ring gear R3. The restraint condition at this time is that the first ring gear R1 and the third ring gear R3 maintain the same rotation speed via the third clutch C3, and the first sun gear S1 and the third ring gear 3 via the first clutch C1. The condition is that the sun gear S3 maintains the same rotational speed. Due to this rotational constraint, the rotation of the third sun gear S3 is reverse to the input rotation (forward rotation) to the first carrier PC1. The rotation of the third sun gear S3 is directly input to the second sun gear S2 via the second rotating member M2. For this reason, in the second planetary gear PG2 fixed to the carrier, the rotation (reverse rotation) input to the second sun gear S2 is further reversely rotated to be forward rotated and output from the second ring gear R2. This output rotational speed (decelerated rotational speed of the input rotational speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the first speed gear stage is achieved.

(2nd gear)
At the second speed (2nd), the third clutch C3, the first brake B1, and the second brake B2 are simultaneously engaged as shown by hatching in FIG.

  By engaging the third clutch C3, the first ring gear R1 and the third ring gear R3 are directly connected. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC. The second carrier PC2 and the third carrier PC3 are fixed to the transmission case TC by the engagement of the second brake B2 and the first rotating member M1.

  Therefore, when the input rotational speed is input to the first carrier PC1 after passing through the input shaft IN, the first planetary gear PG1 fixed with the sun gear decelerates the input rotation to the first carrier PC1 to reduce the first rotation. Output from the ring gear R1. The rotation of the first ring gear R1 passes through the third clutch C3 and is input to the third ring gear R3 as it is. For this reason, in the third planetary gear PG3 fixed to the carrier, the rotation input to the third ring gear R3 is reversed to be the rotation of the third sun gear S3 (the rotation speed in the reverse direction higher than the first speed). . The rotation of the third sun gear S3 passes through the second rotating member M2 and is input to the second sun gear S2 as it is. For this reason, in the second planetary gear PG2 fixed to the carrier, the rotation (reverse rotation) input to the second sun gear S2 is further reversely rotated to be forward rotated and output from the second ring gear R2. This output rotational speed (decelerated rotational speed lower than the input rotational speed and higher than the first speed) is directly transmitted from the second ring gear R2 to the output shaft OUT, and the second speed gear stage is achieved.

(3rd speed)
At the third speed (3rd), the first clutch C1, the third clutch C3, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The first sun gear S1, the second sun gear S2, and the third sun gear S3 are fixed to the transmission case TC by the simultaneous engagement of the first clutch C1 and the first brake B1 and the second rotating member M2. By engaging the third clutch C3, the first ring gear R1 and the third ring gear R3 are directly connected.

  Therefore, when the input rotational speed is input to the first carrier PC1 after passing through the input shaft IN, the first planetary gear PG1 fixed with the sun gear decelerates the input rotation to the first carrier PC1 to reduce the first rotation. Output from the ring gear R1. The rotation of the first ring gear R1 passes through the third clutch C3 and is input to the third ring gear R3 as it is. For this reason, in the third planetary gear PG3 fixed to the sun gear, the rotation input to the third ring gear R3 is decelerated while being forwardly rotated, and is output from the third carrier PC3. The rotation of the third carrier PC3 is input as it is to the second carrier PC2 after passing through the first rotating member M1. For this reason, in the second planetary gear PG2 fixed to the sun gear, the rotation input to the second carrier PC2 is increased in the positive rotation and output from the second ring gear R2. This output rotational speed (decelerated rotational speed lower than the input rotational speed and higher than the second speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the third speed gear stage is achieved.

(4th gear)
At the fourth speed (4th), the second clutch C2, 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 second clutch C2 and the second rotating member M2. By engaging the third clutch C3, the first ring gear R1 and the third ring gear R3 are directly connected. The three rotating elements S3, PC3, R3 of the third planetary gear PG3 are integrally rotated by directly connecting the two rotating elements S3, R3 via the engaged second clutch C2 and third clutch C3. State. Further, the three rotating elements S2, PC2, R2 of the second planetary gear PG2 are integrated by directly connecting the two rotating elements S2, PC2 via the third planetary gear PG3 and both rotating members M1, M2. Rotated state. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input to the first carrier PC1 after passing through the input shaft IN, the first planetary gear PG1 fixed with the sun gear decelerates the input rotation to the first carrier PC1 to reduce the first rotation. Output from the ring gear R1. The rotation of the first ring gear R1 is input as it is to the third sun gear S3 after passing through the second clutch C2, and is input as it is to the third ring gear R3 after passing through the third clutch C3. Therefore, the three rotating elements S3, PC3, R3 of the third planetary gear PG3 rotate together at the same rotational speed as that of the first ring gear R1, and the three rotating elements of the second planetary gear PG2. S2, PC2, and R2 rotate together at the same rotational speed as that of the first ring gear R1. For this reason, the second ring gear R2 outputs a rotation at the same rotational speed as the first ring gear R1, and this output rotational speed (a reduction rotational speed lower than the input rotational speed and higher than the third speed) is the second planetary gear. Transmission from the second ring gear R2 of the gear PG2 to the output shaft OUT is performed as it is, and the fourth speed is achieved.

(5th speed)
At the fifth speed (5th), the third clutch C3, the fourth clutch C4, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By engaging the third clutch C3, the first ring gear R1 and the third ring gear R3 are directly connected. The input shaft IN, the first carrier PC1, the second carrier PC2, and the third carrier PC3 are directly connected by the engagement of the fourth clutch C4 and the first rotating member M1. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input to the first carrier PC1 after passing through the input shaft IN, the first planetary gear PG1 fixed with the sun gear decelerates the input rotation to the first carrier PC1 to reduce the first rotation. Output from the ring gear R1. The rotation of the first ring gear R1 passes through the third clutch C3 and is input to the third ring gear R3 as it is. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotational speed input to the third ring gear R3 and the rotational speed input to the third carrier PC3 (= input rotational speed) are defined. As a result, the rotation speed (rotation speed obtained by increasing the input rotation speed) output from the remaining third sun gear S3 is determined. The rotation of the third sun gear S3 passes through the second rotating member M2 and is input to the second sun gear S2 as it is. For this reason, in the second planetary gear PG2 having two inputs and one output, the rotational speed input to the second sun gear S2 and the rotational speed input to the second carrier PC2 (= input rotational speed) are defined. Thus, the rotation speed output from the remaining second ring gear R2 is determined. This output rotational speed (decelerated rotational speed lower than the input rotational speed and higher than the fourth speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the fifth speed is achieved.

(Sixth speed gear)
At the sixth speed (6th), the second clutch C2, the third clutch C3, and the fourth clutch C4 are simultaneously engaged as shown by hatching in FIG.

  With the simultaneous engagement of the second clutch C2, the third clutch C3 and the fourth clutch C4 and the rotating members M1 and M2, the first planetary gear PG1, the second planetary gear PG2 and the third planetary gear PG3 are respectively obtained. Two rotating elements are directly connected. For this reason, the three rotating elements S1, PC1, R1 of the first planetary gear PG1, the three rotating elements S2, PC2, R2 of the second planetary gear PG2, and the three rotating elements of the third planetary gear PG3. S3, PC3, and R3 are in an integrally rotated state.

  Therefore, when the input rotation speed is input to the first carrier PC1 through the input shaft IN, the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are integrated according to the input rotation speed. Therefore, the rotational speed of the output shaft OUT becomes the same as the input rotational speed from the input shaft IN, and the sixth gear (directly coupled gear) with a gear ratio of 1 is achieved.

(7th speed)
At the seventh speed (7th), the second clutch C2, the fourth clutch C4, 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 second clutch C2 and the second rotating member M2. The input shaft IN, the first carrier PC1, the second carrier PC2, and the third carrier PC3 are directly connected by the engagement of the fourth clutch C4 and the first rotating member M1. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input to the first carrier PC1 after passing through the input shaft IN, the first planetary gear PG1 fixed with the sun gear decelerates the input rotation to the first carrier PC1 to reduce the first rotation. Output from the ring gear R1. The rotation of the first ring gear R1 passes through the second clutch C2 and is directly input to the second sun gear S2. For this reason, in the second planetary gear PG2 having two inputs and one output, the rotational speed input to the second sun gear S2 and the rotational speed input to the second carrier PC2 (= input rotational speed) are defined. Thus, the rotation speed output from the remaining second ring gear R2 is determined. This output rotation speed (acceleration speed higher than the input rotation speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the seventh speed is achieved.

(8th speed)
At the eighth speed (8th), the first clutch C1, the fourth clutch C4, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The first sun gear S1, the second sun gear S2, and the third sun gear S3 are fixed to the transmission case TC by the simultaneous engagement of the first clutch C1 and the first brake B1 and the second rotating member M2. The input shaft IN, the first carrier PC1, the second carrier PC2, and the third carrier PC3 are directly connected by the engagement of the fourth clutch C4 and the first rotating member M1.

  Therefore, when the input rotational speed is input to the second carrier PC2 after passing through the input shaft IN and the fourth clutch C4, the input rotation to the second carrier PC2 is performed in the second planetary gear PG2 fixed with the sun gear. The speed is increased and output from the second ring gear R2. This output rotational speed (accelerated rotational speed higher than the input rotational speed and higher than the seventh speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the eighth speed gear stage is achieved.

(Reverse speed shift stage)
At the reverse speed (Rev), the second clutch C2, the first brake B1, and the second brake B2 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 second clutch C2 and the second rotating member M2. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC. The second carrier PC2 and the third carrier PC3 are fixed to the transmission case TC by the engagement of the second brake B2 and the first rotating member M1.

  Therefore, when the input rotational speed is input to the first carrier PC1 after passing through the input shaft IN, the first planetary gear PG1 fixed with the sun gear decelerates the input rotation to the first carrier PC1 to reduce the first rotation. Output from the ring gear R1. The rotation of the first ring gear R1 passes through the second clutch C2 and is directly input to the second sun gear S2. Therefore, in the second planetary gear PG2 fixed to the carrier, the rotation (forward rotation) input to the second sun gear S2 is reversed and output from the second ring gear R2. This output rotational speed (the rotational speed in the direction opposite to the input rotational direction) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the 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. FIG. 15 is a view showing a gear meshing frequency table at each of the eight forward speeds 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.

First, when comparing the automatic transmission of the first embodiment (FIGS. 1 and 2) and the conventional automatic transmission (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)
The automatic transmission of the conventional example and the automatic transmission of the first embodiment both achieve a forward speed of 8 speeds and a reverse speed of 1 speed by 3 planetary and 6 friction elements. That is, it can be said that the basic configuration and the speed change performance of both automatic transmissions are equivalent.
(Transmission control performance)
Both the automatic transmission of the conventional example and the automatic transmission of the first embodiment achieve the shift to the adjacent gear stage by the single overhanging shift of releasing one friction element and fastening one friction element. Yes. That is, it can be said that the shift control performance (ease of shift control to adjacent shift stages) of the two automatic transmissions is equivalent.

  However, the automatic transmission of the first embodiment has the following “(a) three planetary gears”, “(b) friction loss at each gear stage”, “(c) gear ratio width”, “(d) jumping speed change”. ”Shift control performance” and “(e) reverse power performance” have advantages over the conventional automatic transmission. Each item will be described below.

(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 (a combination of one double pinion type planetary gear and one single pinion type planetary gear).

  In other words, the use of two double-pinion type planetary gears means that the number of gear meshing is large and the transmission efficiency and gear noise of the gear are poor, the pinion gear diameter is small, the durability reliability is lowered, the number of parts is large, and the cost is high. There is a problem of being up.

  On the other hand, in the automatic transmission of the first embodiment, only the first planetary gear PG1 is a double pinion type planetary gear, and the second planetary gear PG2 and the third planetary gear PG3 are planetary gears using a single pinion. Used. As described above, since only one double pinion type planetary gear is used, the number of gear meshes is reduced, the pinion diameter is increased, and the number of parts is reduced compared to the conventional example using two double pinion type planetary gears. Is advantageous.

-Reduction of the number of gear meshes In the case of the automatic transmission of the first embodiment, the number of gear meshes for torque transmission does not decrease from the conventional example, but the same number of gear meshes as the conventional example can be obtained.
That is, one set of double pinion type planetary gears has three meshing times, whereas one set of single pinion type planetary gears has two times of meshing because there is no meshing between pinions. Therefore, in the case of the first embodiment, as shown in FIG. 3, the average number of meshes between the first speed to the eighth speed is 4.75. On the other hand, in the case of the conventional example using two sets of double pinion type planetary gears, as shown in FIG. 15, the average number of meshes between the first speed to the eighth speed is 4.75. That is, in the case of Example 1, although it does not reach a reduction from the average number of meshes of the conventional example, the same number of meshes is achieved.

-Increasing pinion diameter In the case of the automatic transmission according to the first embodiment, the gear diameter of the pinion meshing with the sun gear and the ring gear is increased, so that the durability reliability is improved.
That is, in the case of a single pinion type planetary gear, in order to mesh one pinion with the sun gear and the ring gear, a plurality of pinions having a gear diameter as a 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 type planetary gear, a pair of pinions mesh with each other and mesh with the sun gear and the ring gear, respectively. Therefore, it is necessary to make the diameter smaller than the distance between the two gears. As described above, in the case of a single pinion type planetary gear, the pinion has a larger gear diameter than the double pinion type planetary gear, so that the rigidity and tooth surface strength of the pinion can be increased, and the durability reliability is improved.

-Reduction of the number of parts In the case of the automatic transmission of the first embodiment, the number of parts constituting the planetary gear 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. In the case of a single pinion type planetary gear, the number of pinion shafts is reduced, and the carrier configuration is simplified. As a result, cost reduction can be achieved.

(b) When each friction speed is obtained by fastening friction loss friction elements at each gear, friction loss cannot be avoided due to oil dragging, etc. generated by the slipping friction element (release element). Is preferably as the friction loss is smaller.

  In the case of the conventional automatic transmission, in order to achieve each of the eight forward speeds, two friction elements are simultaneously engaged at each speed as shown in FIG. For this reason, for example, the friction element that idles at the first speed includes four friction elements that idle at each shift speed, such as the second clutch C2, the third clutch C3, the fourth clutch C4, and the first brake B1. Become. For this reason, friction loss due to oil dragging or the like in the four friction elements that rotate idly increases, leading to deterioration in drive energy transmission efficiency. For example, when the conventional automatic transmission is applied to an engine vehicle, friction loss due to four friction elements that idle causes a deterioration in fuel consumption performance.

  On the other hand, in the case of the automatic transmission of the first embodiment, in order to achieve each of the eight forward speeds, as shown in FIG. 2, three friction elements are simultaneously engaged at each speed. . Therefore, for example, the number of friction elements that idle at the first speed stage is three, such as the second clutch C2, the fourth clutch C4, and the first brake B1, at each speed stage. Therefore, compared to the conventional example, the friction loss at the friction element that idles can be reduced, and the drive energy transmission efficiency can be improved. For example, when the automatic transmission according to the first embodiment is applied to an engine vehicle, fuel efficiency can be improved.

(c) Gear ratio width The change ratio of 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 gear ratio of the first planetary gear PG1 is ρ1 = −0.381, the gear ratio of the second planetary gear PG2 is ρ2 = 0.354, When the gear ratio of the third planetary gear PG3 is set to ρ3 = 0.677, RC = 6.53 (= 4.828 / 0.739) is obtained while maintaining an appropriate gear ratio between adjacent gears.

  In other words, while maintaining an appropriate gear ratio, the RC value can be made slightly larger than the conventional example, and both the starting performance at the lowest gear ratio and the high speed fuel efficiency at the highest gear ratio can be achieved. Can be planned. 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.

(d) Shift control performance of jump gear shift For example, as a driving scene on a highway, etc., when the gear stage is traveling at a constant speed at the seventh speed or the eighth speed on the overdrive side, it is intended to overtake the vehicle ahead of the host vehicle When the accelerator depressing operation is performed, a two-step jump down shift from the seventh speed to the fourth speed or a two-step jump down shift from the eighth speed to the fifth speed is performed.

In the case of the automatic transmission of the conventional example, the two-step jump down shift from the eighth speed to the fifth speed can be achieved by a single overhanging shift of releasing one friction element and fastening one friction element. it can. However, the two-step jump-down shift from the seventh speed to the fourth speed is a double-overlay shift in which two friction elements are released and two friction elements are engaged. Therefore, if there is a shift request from the seventh speed to the fourth speed, the seventh speed → the fifth speed → the fourth speed, or the seventh speed → the sixth speed → the fourth speed, or the seventh speed → It is a speed change that passes through the intermediate speed, such as 6th speed → 5th speed → 4th speed.
Therefore, when the accelerator depression operation is performed with the intention of intermediate acceleration, it takes time from the start of the shift based on the shift command to the end of the shift, and the response of the driving force increase to the intermediate acceleration request that appears in the accelerator depression operation of the driver is delayed. .

On the other hand, in the automatic transmission according to the first embodiment, any shift pattern is selected from the two-step jump down shift from the eighth speed to the fifth speed and the two-step jump down shift from the seventh speed to the fourth speed. Even if the shift command is output, as shown in FIG. 2, it can be achieved by a single overhanging shift of releasing one friction element and fastening one friction element. For this reason, for example, even if there is a shift request from the seventh speed to the fourth speed, the second speed from the seventh speed to the fourth speed is not changed without passing through the intermediate shift speeds (sixth speed, fifth speed). A step-down downshift can be performed.
Therefore, when an accelerator depression operation is performed with the intention of intermediate acceleration, the shift is completed in a short time from the start of the shift based on the shift command, and a response to an increase in driving force with respect to the intermediate acceleration request that appears in the driver's accelerator depression operation is secured. Is done.

(e) Reverse power performance 1st gear ratio and reverse gear ratio are values that determine start acceleration and climbing performance. For example, if the ratio of 1st gear ratio and reverse gear ratio is not near 1, 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 automatic transmission of the conventional example, as shown in FIG. 14, when Rev1 / 1st = 0.750, Rev2 / 1st = 0.469, and selecting the first reverse speed (Rev1) or the second reverse speed (Rev2), The ratio between the first gear ratio and the reverse gear ratio is a value lower than 1, which causes a difference in driving force when switching between forward and reverse travel, resulting in poor reverse start performance.

  On the other hand, in the case of the automatic transmission of the first embodiment, as shown in FIG. 2, Rev / 1st = 0.945, and the ratio of the first gear ratio and the reverse gear ratio is in the vicinity of one. For this reason, there is no difference in driving force when switching between forward and backward travel, and there is no shortage of driving force when starting backward. That is, the vehicle can be operated without impairing the start acceleration performance and the climbing performance.

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, a first carrier PC1 that supports first double pinions P1s and P1r 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 single pinion P2 that meshes with the second sun gear S2 and the second ring gear R2. A second planetary gear PG2, 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. And a third planetary gear PG3 comprising six carriers, and six friction elements. By appropriately fastening and releasing the six friction elements, the speed is changed to at least a forward 8-speed gear stage. Torque from input shaft IN to output shaft OUT In the automatic transmission capable of output, the input shaft IN is always connected to the first carrier PC1, the output shaft OUT is always connected to the second ring gear R2, and the second The carrier PC2 and the third carrier PC3 are always connected to form the first rotating member M1, and the second sun gear S2 and the third sun gear S3 are always connected to perform the second rotation. A first friction element (first clutch C1) that selectively connects between the first sun gear S1 and the second rotating member M2; A second friction element (second clutch C2) that selectively connects the first ring gear R1 and the second rotating member M2, and the first ring gear R1 and the third ring gear R3. A third friction element (third clutch C3) for selectively connecting the first clutch and the first clutch. A fourth friction element (fourth clutch C4) that selectively connects the rear PC1 and the first rotation member M1 and a fifth friction element that can lock the rotation of the first sun gear S1 ( A first brake B1) and a sixth friction element (second brake B2) capable of locking the rotation of the first rotation member M1, and among the six friction elements, three simultaneous engagements are made. By combining the above, at least 8 forward speeds and 1 reverse speed are achieved.
Therefore, while achieving 8 forward speeds with 3 planetary 6 friction elements, it is advantageous in terms of gear efficiency, gear noise, durability reliability, and cost, and also improves drive energy transmission efficiency by minimizing friction loss. Can be achieved.

(2) Of the six friction elements, the first friction element (first clutch C1) and the third friction element (third clutch C3) are at least in the eighth forward speed by the combination of three simultaneous engagements. The first speed achieved by simultaneous engagement of the sixth friction element (second brake B2), the third friction element (third clutch C3), and the fifth friction element (first brake B1). A second speed achieved by simultaneous engagement of the sixth friction element (second brake B2), the first friction element (first clutch C1), the third friction element (third clutch C3), and the Third speed achieved by simultaneous engagement of the fifth friction element (first brake B1), the second friction element (second clutch C2), the third friction element (third clutch C3), and the second 4th speed achieved by simultaneous engagement of 5 friction elements (first brake B1), A third speed achieved by simultaneous engagement of the third friction element (third clutch C3), the fourth friction element (fourth clutch C4), and the fifth friction element (first brake B1), and the second A sixth speed achieved by simultaneous engagement of the second friction element (second clutch C2), the third friction element (third clutch C3), and the fourth friction element (fourth clutch C4), and the second The seventh speed achieved by simultaneous engagement of the friction element (second clutch C2), the fourth friction element (fourth clutch C4), and the fifth friction element (first brake B1), and the first friction And an eighth speed achieved by simultaneous engagement of the element (first clutch C1), the fourth friction element (fourth clutch C4), and the fifth friction element (first brake B1).
For this reason, the shift to the adjacent stage is achieved by single replacement by fastening one friction element and releasing one friction element, and the shift control is simplified and advantageous, and an appropriate interstage ratio is achieved. While maintaining the RC value, the RC value can be set to a required value for achieving both the starting performance at the lowest gear ratio and the high speed fuel consumption at the highest gear ratio. In addition, even if a shift command is output according to any shift pattern among the two-step jump shift between the eighth speed and the fifth speed and the two-step jump shift between the seventh speed and the fourth speed, 1 Downshifting and upshifting can be achieved by heavy switching, and it is possible to respond to acceleration requests and deceleration requests with good response.

(3) Of the six friction elements, the first reverse speed achieved by the combination of three simultaneous engagements is the second friction element (second clutch C2) and the fifth friction element (first brake B1). And the sixth friction element (second brake B2).
For this reason, even if a gear ratio that achieves an appropriate RC value and interstage ratio is selected, the value for evaluating the reverse gear ratio (= reverse gear ratio / 1st gear ratio) can be close to 1. By ensuring the driving force equivalent to that when starting forward as the driving force when starting backward, it is possible to prevent the driver from feeling uncomfortable when repeatedly moving forward and backward, such as when entering the garage, and to start backward. Sometimes it is possible to prevent the driving force from becoming insufficient.

The second embodiment is a combination of one double pinion type planetary gear and two single pinion type planetary gears as in the first embodiment, but the third planetary gear is a double pinion type planetary gear.

First, the configuration will be described.
FIG. 16 is a skeleton diagram illustrating the automatic transmission according to the second embodiment. Hereinafter, the planetary gear configuration and the friction element configuration of the automatic transmission according to the second embodiment will be described with reference to FIG.

  As shown in FIG. 16, the automatic transmission according to the second 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, The first rotary member M1, the second rotary member M2, the first clutch C1 (first friction element), the second clutch C2 (second friction element), and the third clutch C3 (third Friction element), fourth clutch C4 (fourth friction element), first brake B1 (fifth friction element), second brake B2 (sixth friction element), and transmission case TC. I have.

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

  The second planetary gear PG2 is a single pinion type planetary gear, and a second single gear P2 meshing with the second sun gear S2, the second ring gear R2, the second sun gear S2 and the second ring gear R2. And a second carrier PC2 that supports the second carrier PC2.

  The third planetary gear PG3 is a double pinion type planetary gear, and a third sun gear S3, a third ring gear R3, and a third double pinion P3s meshing with the third sun gear S3 and the third ring gear R3. , P3r and a third carrier PC3.

  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 ring gear R1.

  The output shaft OUT is a shaft for outputting the rotational drive torque after shifting to the drive wheels via a propeller shaft, final gear, or the like, and is always connected to the second ring gear R2.

  The first rotating member M1 is a rotating member that always connects the second carrier PC2 and the third ring gear R3 without interposing a friction element.

  The second rotating member M2 is a rotating member that always connects the second sun gear S2 and the third sun gear S3 without interposing a friction element.

  The first clutch C1 is a first friction element that selectively connects the first sun gear S1 and the second rotating member M2.

  The second clutch C2 is a second friction element that selectively connects the first carrier PC1 and the second rotating member M2.

  The third clutch C3 is a third friction element that selectively connects the first carrier PC1 and the third carrier PC3.

  The fourth clutch C4 is a fourth friction element that selectively connects the first ring gear R1 and the first rotating member M1.

  The first brake B1 is a fifth friction element that can lock the rotation of the first sun gear S1 with respect to the transmission case TC.

  The second brake B2 is a sixth friction element that can lock the rotation of the first rotating member M1 with respect to the transmission case TC.

  As shown in FIG. 16, 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 in order.

  FIG. 17 is a diagram illustrating a fastening operation table that achieves eight forward speeds and one reverse speed by combining three simultaneous fastenings among six friction elements in the automatic transmission according to the second embodiment. FIG. 18 is a diagram illustrating a gear meshing frequency table at each of the eight forward speeds in the automatic transmission according to the second embodiment. Hereinafter, based on FIG.17 and FIG.18, the transmission structure which establishes each gear stage of the automatic transmission of Example 2 is demonstrated.

  The automatic transmission according to the second embodiment has eight forward speeds and one reverse speed as described below by combining three of the six friction elements C1, C2, C3, C4, B1, and B2 simultaneously. To achieve.

  As shown in FIG. 17, the first speed (1st) is achieved by simultaneously engaging the first clutch C1, the third clutch C3, and the second brake B2. As shown in FIG. 18, the number of gear meshes at the first 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 + 2 times + 3 times).

  As shown in FIG. 17, the second speed (2nd) is achieved by simultaneously engaging the third clutch C3, the first brake B1, and the second brake B2. As shown in FIG. 18, since the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are involved in the meshing, the number of gear meshes at the second speed gear stage is the total number of gear meshes. Is 7 times (= 2 times + 2 times + 3 times).

  As shown in FIG. 17, the third speed (3rd) is achieved by simultaneously engaging the first clutch C1, the third clutch C3, and the first brake B1. As shown in FIG. 18, since the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are involved in the meshing, the number of gear meshes at the third speed is the total number of gear meshes. Is 7 times (= 2 times + 2 times + 3 times).

  As shown in FIG. 17, the fourth speed (4th) is achieved by simultaneously engaging the second clutch C2, the third clutch C3, and the first brake B1. As shown in FIG. 18, the number of gear meshes at the fourth speed gear stage is only 2nd (= 2 times + 0 times + 0 times) because only the first planetary gear PG1 is involved in meshing. It becomes.

  As shown in FIG. 17, the fifth speed (5th) is achieved by simultaneously engaging the third clutch C3, the fourth clutch C4, and the first brake B1. As shown in FIG. 18, the number of gear meshes at the fifth 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 + 2 times + 3 times).

  As shown in FIG. 17, the sixth speed (6th) is achieved by simultaneously engaging the second clutch C2, the third clutch C3, and the fourth clutch C4. As shown in FIG. 18, the number of gear meshes at the sixth 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. 17, the seventh speed (7th) is achieved by simultaneously engaging the second clutch C2, the fourth clutch C4, and the first brake B1. As shown in FIG. 18, since the first planetary gear PG1 and the second planetary gear PG2 are involved in meshing, the total number of gear engagements at the seventh speed is 4 (= 2). Times +2 times +0 times).

  As shown in FIG. 17, the eighth speed (8th) is established by simultaneously engaging the first clutch C1, the fourth clutch C4, and the first brake B1. As shown in FIG. 18, the number of gear meshes at the eighth speed gear stage is only 2nd planetary gear PG2 is involved in meshing, so the total number of times is 2 times (= 0 times + 2 times + 0 times). It becomes.

  As shown in FIG. 17, the reverse speed (Rev) is achieved by simultaneously engaging the second clutch C2, the first brake B1, and the second brake B2.

Next, the operation will be described.
The operation of the automatic transmission according to the second embodiment will be described by dividing it into “shift operation at each gear 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, the third clutch C3, and the second brake B2 are simultaneously engaged as shown by hatching in FIG.

  The first sun gear S1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the first clutch C1 and the second rotating member M2. By engaging the third clutch C3, the first carrier PC1 and the third carrier PC3 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 second brake B2 and the first rotating member M1.

  Therefore, when the input rotational speed is input after passing through the input shaft IN to the first ring gear R1, the rotation of the first sun gear S1 of the first planetary gear PG1 and the first carrier PC1 causes the first rotation of the ring gear fixed. The third planetary gear PG3 rotates while being restricted by the rotation of the third sun gear S3 and the third carrier PC3. The constraint condition at this time is that the first carrier PC1 and the third carrier PC3 maintain the same rotational speed via the third clutch C3, and the first sun gear S1 and the third carrier PC3 via the first clutch C1. The condition is that the sun gear S3 maintains the same rotational speed. Due to this rotational constraint, the rotation of the third sun gear S3 is the reverse of the input rotation (forward rotation) to the first ring gear R1. The rotation of the third sun gear S3 is directly input to the second sun gear S2 via the second rotating member M2. For this reason, in the second planetary gear PG2 fixed to the carrier, the rotation (reverse rotation) input to the second sun gear S2 is further reversely rotated to be forward rotated and output from the second ring gear R2. This output rotational speed (decelerated rotational speed of the input rotational speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the first speed gear stage is achieved.

(2nd gear)
At the second speed (2nd), the third clutch C3, the first brake B1, and the second brake B2 are simultaneously engaged as shown by hatching in FIG.

  By engaging the third clutch C3, the first carrier PC1 and the third carrier PC3 are directly connected. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC. The second carrier PC2 and the third ring gear R3 are fixed to the transmission case TC by the engagement of the second brake B2 and the first rotating member M1.

  Therefore, when the input rotational speed is input after passing through the input shaft IN to the first ring gear R1, the first planetary gear PG1 fixed to the sun gear decelerates the input rotation to the first ring gear R1 and the first Output from the carrier PC1. The rotation of the first carrier PC1 is input as it is to the third carrier PC3 after passing through the third clutch C3. Therefore, in the third planetary gear PG3 fixed to the ring gear, the rotation input to the third carrier PC3 is reversed to rotate the third sun gear S3 (reverse rotation at a higher rotational speed than the first speed). To do. The rotation of the third sun gear S3 passes through the second rotating member M2 and is input to the second sun gear S2 as it is. For this reason, in the second planetary gear PG2 fixed to the carrier, the rotation (reverse rotation) input to the second sun gear S2 is further reversely rotated to be forward rotated and output from the second ring gear R2. This output rotational speed (decelerated rotational speed lower than the input rotational speed and higher than the first speed) is directly transmitted from the second ring gear R2 to the output shaft OUT, and the second speed gear stage is achieved.

(3rd speed)
At the third speed (3rd), the first clutch C1, the third clutch C3, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The first sun gear S1, the second sun gear S2, and the third sun gear S3 are fixed to the transmission case TC by the simultaneous engagement of the first clutch C1 and the first brake B1 and the second rotating member M2. By engaging the third clutch C3, the first carrier PC1 and the third carrier PC3 are directly connected.

  Therefore, when the input rotational speed is input after passing through the input shaft IN to the first ring gear R1, the first planetary gear PG1 fixed to the sun gear decelerates the input rotation to the first ring gear R1 and the first Output from the carrier PC1. The rotation of the first carrier PC1 passes through the third clutch C3 and is directly input to the third carrier PC3. For this reason, in the third planetary gear PG3 fixed to the sun gear, the rotation input to the third carrier PC3 is decelerated while being forwardly rotated, and is output from the third ring gear R3. The rotation of the third ring gear R3 passes through the first rotating member M1 and is input as it is to the second carrier PC2. For this reason, in the second planetary gear PG2 fixed to the sun gear, the rotation input to the second carrier PC2 is increased in the positive rotation and output from the second ring gear R2. This output rotational speed (decelerated rotational speed lower than the input rotational speed and higher than the second speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the third speed gear stage is achieved.

(4th gear)
At the fourth speed (4th), the second clutch C2, the third clutch C3, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The first carrier PC1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the second clutch C2 and the second rotating member M2. By engaging the third clutch C3, the first carrier PC1 and the third carrier PC3 are directly connected. The three rotating elements S3, PC3, R3 of the third planetary gear PG3 are integrally rotated by directly connecting the two rotating elements S3, PC3 via the engaged second clutch C2 and third clutch C3. State. Further, the three rotating elements S2, PC2, R2 of the second planetary gear PG2 are integrated by directly connecting the two rotating elements S2, PC2 via the third planetary gear PG3 and both rotating members M1, M2. Rotated state. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input after passing through the input shaft IN to the first ring gear R1, the first planetary gear PG1 fixed to the sun gear decelerates the input rotation to the first ring gear R1 and the first Output from the carrier PC1. The rotation of the first carrier PC1 is input as it is to the third sun gear S3 after passing through the second clutch C2, and is input as it is to the third carrier PC3 after passing through the third clutch C3. Accordingly, the three rotating elements S3, PC3, and R3 of the third planetary gear PG3 rotate together at the same rotational speed as that of the first carrier PC1, and the three rotating elements of the second planetary gear PG2. S2, PC2, and R2 rotate together at the same rotational speed as that of the first carrier PC1. Therefore, the second ring gear R2 outputs a rotation at the same rotation speed as that of the first carrier PC1, and this output rotation speed (a reduction rotation speed lower than the input rotation speed and higher than the third speed) is the second planetary gear. Transmission from the second ring gear R2 of the gear PG2 to the output shaft OUT is performed as it is, and the fourth speed is achieved.

(5th speed)
At the fifth speed (5th), the third clutch C3, the fourth clutch C4, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By engaging the third clutch C3, the first carrier PC1 and the third carrier PC3 are directly connected. The input shaft IN, 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 first rotating member M1. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input after passing through the input shaft IN to the first ring gear R1, the first planetary gear PG1 fixed to the sun gear decelerates the input rotation to the first ring gear R1 and the first Output from the carrier PC1. The rotation of the first carrier PC1 passes through the third clutch C3 and is directly input to the third carrier PC3. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotational speed input to the third carrier PC3 and the rotational speed input to the third ring gear R3 (= input rotational speed) are defined. Thus, the output rotational speed from the remaining third sun gear S3 (the rotational speed obtained by increasing the input rotational speed while maintaining the positive rotational speed) is determined. The rotation of the third sun gear S3 passes through the second rotating member M2 and is input to the second sun gear S2 as it is. For this reason, in the second planetary gear PG2 having two inputs and one output, the rotational speed input to the second sun gear S2 and the rotational speed input to the second carrier PC2 (= input rotational speed) are defined. Thus, the output rotational speed from the remaining second ring gear R2 is determined. This output rotational speed (decelerated rotational speed lower than the input rotational speed and higher than the fourth speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the fifth speed is achieved.

(Sixth speed gear)
At the sixth speed (6th), the second clutch C2, the third clutch C3, and the fourth clutch C4 are simultaneously engaged as shown by hatching in FIG.

  With the simultaneous engagement of the second clutch C2, the third clutch C3 and the fourth clutch C4 and the rotating members M1 and M2, the first planetary gear PG1, the second planetary gear PG2 and the third planetary gear PG3 are respectively obtained. Two rotating elements are directly connected. For this reason, the three rotating elements S1, PC1, R1 of the first planetary gear PG1, the three rotating elements S2, PC2, R2 of the second planetary gear PG2, and the three rotating elements of the third planetary gear PG3. S3, PC3, and R3 are in an integrally rotated state.

  Therefore, when the input rotational speed is input after passing through the input shaft IN to the first ring gear R1, the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are integrated according to the input rotational speed. Therefore, the rotational speed of the output shaft OUT becomes the same as the input rotational speed from the input shaft IN, and the sixth gear (directly coupled gear) with a gear ratio of 1 is achieved.

(7th speed)
At the seventh speed (7th), the second clutch C2, the fourth clutch C4, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The first carrier PC1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the second clutch C2 and the second rotating member M2. The input shaft IN, 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 first rotating member M1. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input after passing through the input shaft IN to the first ring gear R1, the first planetary gear PG1 fixed to the sun gear decelerates the input rotation to the first ring gear R1 and the first Output from the carrier PC1. The rotation of the first carrier PC1 passes through the second clutch C2 and is directly input to the second sun gear S2. For this reason, in the second planetary gear PG2 having two inputs and one output, the rotational speed input to the second sun gear S2 and the rotational speed input to the second carrier PC2 (= input rotational speed) are defined. Thus, the output rotational speed from the remaining second ring gear R2 is determined. This output rotation speed (acceleration speed higher than the input rotation speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the seventh speed is achieved.

(8th speed)
At the eighth speed (8th), the first clutch C1, the fourth clutch C4, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The first sun gear S1, the second sun gear S2, and the third sun gear S3 are fixed to the transmission case TC by the simultaneous engagement of the first clutch C1 and the first brake B1 and the second rotating member M2. The input shaft IN, 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 first rotating member M1.

  Therefore, when the input rotational speed is input to the second carrier PC2 after passing through the input shaft IN and the fourth clutch C4, the input rotation to the second carrier PC2 is performed in the second planetary gear PG2 fixed with the sun gear. The speed is increased and output from the second ring gear R2. This output rotational speed (accelerated rotational speed higher than the input rotational speed and higher than the seventh speed) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and the eighth speed gear stage is achieved.

(Reverse speed shift stage)
At the reverse speed (Rev), the second clutch C2, the first brake B1, and the second brake B2 are simultaneously engaged as shown by hatching in FIG.

  The first carrier PC1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the second clutch C2 and the second rotating member M2. By engaging the first brake B1, the first sun gear S1 is fixed to the transmission case TC. The second carrier PC2 and the third ring gear R3 are fixed to the transmission case TC by the engagement of the second brake B2 and the first rotating member M1.

  Therefore, when the input rotational speed is input after passing through the input shaft IN to the first ring gear R1, the first planetary gear PG1 fixed to the sun gear decelerates the input rotation to the first ring gear R1 and the first Output from the carrier PC1. The rotation of the first carrier PC1 passes through the second clutch C2 and is directly input to the second sun gear S2. Therefore, in the second planetary gear PG2 fixed to the carrier, the rotation (forward rotation) input to the second sun gear S2 is reversed and output from the second ring gear R2. This output rotation speed (rotation speed in the direction opposite to the input rotation direction) is transmitted as it is from the second ring gear R2 to the output shaft OUT, and a reverse gear is achieved.

[Advantages by comparison with conventional technology]
Hereinafter, the advantages of the automatic transmission of the second embodiment in comparison with the prior art will be described with reference to FIGS. 13 to 15. The automatic transmission of the second embodiment (FIGS. 16 and 17) and the automatic transmission of the conventional example will be described. Comparing the transmissions (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)
The automatic transmission of the conventional example and the automatic transmission of the first embodiment both achieve a forward speed of 8 speeds and a reverse speed of 1 speed by 3 planetary and 6 friction elements. That is, it can be said that the basic configuration and the speed change performance of both automatic transmissions are equivalent.
(Transmission control performance)
Both the automatic transmission of the conventional example and the automatic transmission of the first embodiment achieve the shift to the adjacent gear stage by the single overhanging shift of releasing one friction element and fastening one friction element. Yes. That is, it can be said that the shift control performance (ease of shift control to adjacent shift stages) of the two automatic transmissions is equivalent.

  However, the automatic transmission of the second embodiment has the following “(a) three planetary gears”, “(b) friction loss at each gear stage”, “(c) gear ratio width”, “(d) jumping speed change”. ”Shift control performance” and “(e) reverse power performance” have advantages over the conventional automatic transmission. Each item will be described below.

(a) Three planetary gear In the case of the automatic transmission of the second embodiment, only the third planetary gear PG3 is a double pinion type planetary gear, and the first planetary gear PG1 and the second planetary gear PG2 are single pinion type planetary gears Gears are used. As described above, since only one double pinion type planetary gear is used, the number of gear meshes is reduced, the pinion diameter is increased, and the number of parts is reduced compared to the conventional example using two double pinion type planetary gears. Is advantageous.

Reduction of the number of gear meshes In the case of the automatic transmission of the second embodiment, the number of gear meshes for torque transmission is reduced as compared with the conventional example, the gear transmission efficiency is improved, and the gear noise is reduced.
That is, one set of double pinion type planetary gears has three meshing times, whereas one set of single pinion type planetary gears has two times of meshing because there is no meshing between pinions. Therefore, in the case of the second embodiment, as shown in FIG. 18, the average number of meshes between the first speed to the eighth speed is 4.50. On the other hand, in the case of the conventional example using two sets of double pinion type planetary gears, as shown in FIG. 15, the average number of meshes between the first speed to the eighth speed is 4.75. That is, in the case of Example 2, the average meshing number reduced by 0.25 from the average meshing number of the conventional example is achieved.

-Increasing pinion diameter In the case of the automatic transmission of the second embodiment, the diameter of the pinion meshing with the sun gear and the ring gear is increased, so that the durability reliability is improved.
That is, in the case of a single pinion type planetary gear, in order to mesh one pinion with the sun gear and the ring gear, a plurality of pinions having a gear diameter as a 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 type planetary gear, a pair of pinions mesh with each other and mesh with the sun gear and the ring gear, respectively. Therefore, it is necessary to make the diameter smaller than the distance between the two gears. As described above, in the case of a single pinion type planetary gear, the pinion has a larger gear diameter than the double pinion type planetary gear, so that the rigidity and tooth surface strength of the pinion can be increased, and the durability reliability is improved.

-Reduction in the number of parts In the case of the automatic transmission of the second embodiment, the number of parts constituting the planetary gear 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. In the case of a single pinion type planetary gear, the number of pinion shafts is reduced, and the carrier configuration is simplified. As a result, cost reduction can be achieved.

(b) When each friction speed is obtained by fastening friction loss friction elements at each gear, friction loss cannot be avoided due to oil dragging, etc. generated by the slipping friction element (release element). Is preferably as the friction loss is smaller.

  In the case of the conventional automatic transmission, in order to achieve each of the eight forward speeds, two friction elements are simultaneously engaged at each speed as shown in FIG. For this reason, for example, the friction element that idles at the first speed includes four friction elements that idle at each shift speed, such as the second clutch C2, the third clutch C3, the fourth clutch C4, and the first brake B1. Become. For this reason, friction loss due to oil dragging or the like in the four friction elements that rotate idly increases, leading to deterioration in drive energy transmission efficiency. For example, when the conventional automatic transmission is applied to an engine vehicle, friction loss due to four friction elements that idle causes a deterioration in fuel consumption performance.

  On the other hand, in the case of the automatic transmission of the second embodiment, in order to achieve each of the eight forward speeds, as shown in FIG. 17, three friction elements are simultaneously engaged at each speed. . Therefore, for example, the number of friction elements that idle at the first speed stage is three, such as the second clutch C2, the fourth clutch C4, and the first brake B1, at each speed stage. Therefore, compared to the conventional example, the friction loss at the friction element that idles can be reduced, and the drive energy transmission efficiency can be improved. For example, when the automatic transmission according to the second embodiment is applied to an engine vehicle, fuel efficiency can be improved.

(c) Gear ratio width The change ratio of 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 second embodiment, as shown in FIG. 17, the gear ratio of the first planetary gear PG1 is ρ1 = 0.61, the gear ratio of the second planetary gear PG2 is ρ2 = 0.35, When the gear ratio of the planetary gear PG3 is set to ρ3 = -0.40, RC = 6.53 (= 4.829 / 0.739) is obtained while maintaining an appropriate gear ratio between adjacent gears.

  In other words, while maintaining an appropriate gear ratio, the RC value can be made slightly larger than the conventional example, and both the starting performance at the lowest gear ratio and the high speed fuel efficiency at the highest gear ratio can be achieved. Can be planned. 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.

(d) Shift control performance of jump shift In the case of the automatic transmission of the conventional example, two-step jump down shift from the seventh speed to the fourth speed includes two releases of two friction elements and engagement of two friction elements. Since this is a double-overshift gear shift, when there is a shift request from the seventh speed to the fourth speed, the shift is performed after the intermediate gear stage.
Therefore, when the accelerator depression operation is performed with the intention of intermediate acceleration, it takes time from the start of the shift based on the shift command to the end of the shift, and the response of the driving force increase to the intermediate acceleration request that appears in the accelerator depression operation of the driver is delayed. .

On the other hand, in the automatic transmission of the second embodiment, any shift pattern is selected from the two-step jump down shift from the eighth speed to the fifth speed and the two-step jump down shift from the seventh speed to the fourth speed. Even if a gear shift command is output, it can be achieved by a single overhanging shift of releasing one friction element and fastening one friction element, as shown in FIG. For this reason, for example, even if there is a shift request from the seventh speed to the fourth speed, the second speed from the seventh speed to the fourth speed is not changed without passing through the intermediate shift speeds (sixth speed, fifth speed). A step-down downshift can be performed.
Therefore, when an accelerator depression operation is performed with the intention of intermediate acceleration, the shift is completed in a short time from the start of the shift based on the shift command, and a response to an increase in driving force with respect to the intermediate acceleration request that appears in the driver's accelerator depression operation is secured. Is done.

(e) Reverse power performance 1st gear ratio and reverse gear ratio are values that determine start acceleration and climbing performance. For example, if the ratio of 1st gear ratio and reverse gear ratio is not near 1, 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 automatic transmission of the conventional example, as shown in FIG. 14, when Rev1 / 1st = 0.750, Rev2 / 1st = 0.469, and selecting the first reverse speed (Rev1) or the second reverse speed (Rev2), The ratio between the first gear ratio and the reverse gear ratio is a value lower than 1, which causes a difference in driving force when switching between forward and reverse travel, resulting in poor reverse start performance.

  On the other hand, in the case of the automatic transmission of the second embodiment, as shown in FIG. 17, Rev / 1st = 0.948, and the ratio between the first gear ratio and the reverse gear ratio is in the vicinity of one. For this reason, there is no difference in driving force when switching between forward and backward travel, and there is no shortage of driving force when starting backward. That is, the vehicle can be operated without impairing the start acceleration performance and the climbing performance.

Next, the effect will be described.
In the automatic transmission according to the second embodiment, the following effects can be obtained in addition to the effects (2) and (3) of the first embodiment.

(4) The first sun gear S1, the first ring gear R1, and the first carrier PC1 supporting the first sun gear S1 and the first single pinion P1 meshing with the first ring gear R1. A second carrier that supports a first planetary gear PG1, a second sun gear S2, a second ring gear R2, and a second single pinion P2 meshing with the second sun gear S2 and the second ring gear R2. A second planetary gear PG2 comprising PC2, a third sun gear S3, a third ring gear R3, and third double pinions P3s and P3r meshing with the third sun gear S3 and the third ring gear R3. And a third planetary gear PG3 comprising six carriers, and six friction elements. By appropriately fastening and releasing the six friction elements, the speed is changed to at least a forward 8-speed gear stage. Torque from input shaft IN to output shaft OUT In the automatic transmission capable of output, the input shaft IN is always connected to the first ring gear R1, the output shaft OUT is always connected to the second ring gear R2, and the second The carrier PC2 and the third ring gear R3 are always connected to form the first rotating member M1, and the second sun gear S2 and the third sun gear S3 are always connected to perform the second rotation. A first friction element (first clutch C1) that selectively connects between the first sun gear S1 and the second rotating member M2; A second friction element (second clutch C2) for selectively connecting the first carrier PC1 and the second rotating member M2, and the first carrier PC1 and the third carrier PC3. A third friction element (third clutch C3) for selectively connecting between the first phosphorus and the first phosphorus A fourth friction element (fourth clutch C4) that selectively connects between the gear R1 and the first rotating member M1, and a fifth friction element that can lock the rotation of the first sun gear S1 ( A first brake B1) and a sixth friction element (second brake B2) capable of locking the rotation of the first rotation member M1, and among the six friction elements, three simultaneous engagements are made. By combining the above, at least 8 forward speeds and 1 reverse speed are achieved.
Therefore, while achieving 8 forward speeds with 3 planetary 6 friction elements, it is advantageous in terms of gear efficiency, gear noise, durability reliability, and cost, and also improves drive energy transmission efficiency by minimizing friction loss. Can be achieved.

  As mentioned above, although the automatic transmission of this invention has been demonstrated based on Example 1 and Example 2, it is not restricted to these Examples about a concrete structure, Each claim of a claim is a claim. Design changes and additions are allowed without departing from the gist of the invention.

  In the first embodiment, the gear ratio of the first planetary gear PG1 is ρ1 = −0.381, the gear ratio of the second planetary gear PG2 is ρ2 = 0.354, and the gear ratio of the third planetary gear PG3 is ρ3 = 0.679. In Example 2, the gear ratio of the first planetary gear PG1 is ρ1 = 0.614, the gear ratio of the second planetary gear PG2 is ρ2 = 0.353, and the gear ratio of the third planetary gear PG3 is An example in which ρ3 = -0.403 is shown. However, the gear ratio ρ of each planetary gear PG1, PG2, PG3 is a value within a range in which the gear ratio can be set, and is set so as to obtain a gear ratio with a high RC value and an appropriate interstage ratio. If so, the specific values are not limited to those in the first and second embodiments.

  In the first and second embodiments, an example of an automatic transmission that is applied to an FR engine vehicle in which the input / output shafts are coaxially arranged has been shown. The present invention can also be applied as an automatic transmission for various vehicles such as battery cars.

PG1 first planetary gear
PG2 Second planetary gear
PG3 3rd planetary gear
IN input shaft
OUT output shaft
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 (3)

A first planetary gear comprising: a first sun gear; a first ring gear; a first carrier that supports the first sun gear and a first double 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 single pinion meshing 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 second ring gear,
The second carrier and the third carrier are always connected to form a first rotating member,
The second sun gear and the third sun gear are always connected to form a second rotating member,
The six friction elements are:
A first friction element that selectively connects between the first sun gear and the second rotating member;
A second friction element that selectively connects between the first ring gear and the second rotating member;
A third friction element that selectively connects between the first ring gear and the third ring gear;
A fourth friction element that selectively connects between the first carrier and the first rotating member;
A fifth friction element capable of locking the rotation of the first sun gear;
A sixth friction element capable of locking the rotation of the first rotating member;
Composed of
An automatic transmission characterized in that at least eight forward speeds and one reverse speed are achieved by a combination of three simultaneous engagements among the six friction elements. The automatic transmission according to claim 1 , wherein
Among the six friction elements, at least eight forward speeds are achieved by a combination of three simultaneous fastenings.
A first speed achieved by simultaneous engagement of the first friction element, the third friction element, and the sixth friction element;
A second speed achieved by simultaneous engagement of the third friction element, the fifth friction element and the sixth friction element;
A third speed achieved by simultaneous engagement of the first friction element, the third friction element, and the fifth friction element;
A fourth speed achieved by simultaneous engagement of the second friction element, the third friction element, and the fifth friction element;
A fifth speed achieved by simultaneous engagement of the third friction element, the fourth friction element, and the fifth friction element;
A sixth speed achieved by simultaneous engagement of the second friction element, the third friction element, and the fourth friction element;
A seventh speed achieved by simultaneous engagement of the second friction element, the fourth friction element, and the fifth friction element;
An eighth speed achieved by simultaneous engagement of the first friction element, the fourth friction element, and the fifth friction element;
An automatic transmission characterized by comprising: The automatic transmission according to claim 1 or 2 ,
Of the six friction elements, the first reverse speed achieved by a combination of three simultaneous engagements is achieved by simultaneous engagement of the second friction element, the fifth friction element, and the sixth friction element. Automatic transmission featured.

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