JP2011033137A - Automatic transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve the efficiency of transmission of drive energy with advantages in terms of gear efficiency, gear noise, duration reliability and cost by attaining eight forward speeds by three planetary and six frictional elements. <P>SOLUTION: The automatic transmission includes a single pinion first planetary gearset PG1, a single pinion third planetary gearset PG3, and a double pinion second planetary gearset PG2. An input shaft IN is regularly coupled to a sun gear S1, an output shaft OUT is regularly coupled to a ring gear R3, and a ring gear R1 is regularly coupled to a ring gear R2 to constitute a rotary member M1, and a carrier PC2 is regularly coupled to a sun gear S3 to constitute a rotary member M2. The transmission further includes: a first clutch C1 coupling the sun gear S1 to a carrier PC3, a second clutch C2 coupling a carrier PC1 to the carrier PC3; a third clutch C3 coupling the carrier PC1 to the carrier PC2; a fourth clutch C4 coupling a sun gear S2 to the carrier PC2; a first brake B1 capable of locking the rotation of the sun gear S2; and a second brake B2 capable of locking the rotation of the carrier PC3. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

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

  Conventionally, as an automatic transmission that achieves a forward 8-speed with three planets and six friction elements, a double-pinion planetary gear, a Rabinio planetary gear unit (one double-pinion planet and one single-pinion planet), 4 One having one clutch and two brakes is known (for example, see Patent Document 1).

JP 2001-182785 A

However, although the conventional automatic transmission achieves a forward eight-speed shift stage with three planets and six friction elements, it uses two double-pinion planetary gears, which is disadvantageous for the following items: There was a problem.
-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. While achieving eight forward speeds with three planetary and six friction elements, it has advantages in terms of gear efficiency, gear noise, durability reliability, and cost, and transmission of driving energy. An object of the present invention is to provide an automatic transmission capable of improving efficiency.

In order to achieve the above object, an automatic transmission according to the present invention includes a first sun gear, a first carrier that supports a first pinion that meshes with the first sun gear, and a first carrier that meshes with the first pinion. A first planetary gear comprising a ring gear of
A second planetary gear comprising a second sun gear, a second ring gear, and a second carrier that supports a second double pinion that meshes with the second sun gear and the second ring gear, respectively.
A third planetary gear comprising a third sun gear, a third carrier supporting a third pinion meshing with the third sun gear, and a third ring gear meshing with the third pinion;
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 sun gear,
The output shaft is always connected to the third ring gear,
The first ring gear and the second ring gear are always connected to form a first rotating member,
The second carrier 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 third carrier;
A second friction element that selectively couples between the first carrier and the third carrier;
A third friction element for selectively connecting between the first carrier and the second rotating member;
A fourth friction element that selectively connects between two rotating elements of the second planetary gear;
A fifth friction element capable of locking the rotation of the second sun gear;
A sixth friction element capable of locking the rotation of the third carrier;
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. Among these, for three planets, a first planetary gear by a single pinion and a second planetary gear by a third planetary gear and a double pinion are used. For this reason, compared with the case where the planetary gear by two double pinions is used, the number of gear meshes is reduced, the gear efficiency is improved, and the gear noise is reduced. 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 eight forward speeds with the three planetary six friction elements, it is possible to achieve advantages in terms of gear efficiency, gear noise, durability reliability, and cost, and improvement in drive energy transmission efficiency.

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

  Hereinafter, the best mode for realizing an automatic transmission according to the present invention will be described based on Examples 1 to 3 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 single pinion type planetary gear, and includes a first sun gear S1, a first carrier PC1 that supports the first pinion P1 that meshes with the first sun gear S1, and the first And a first ring gear R1 meshing with the pinion P1.

  The second planetary gear PG2 is a double pinion type planetary gear, and is engaged with the second sun gear S2, the second ring gear R2, the second sun gear S2, and the second ring gear R2, respectively. And a second carrier PC2 supporting P2s and P2r.

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

  The input shaft IN is a shaft to which rotational drive torque from a drive source (engine or the like) is input via a torque converter or the like, and is always connected to the first sun gear S1.

  The output shaft OUT is a shaft that outputs a rotational drive torque after shifting to the drive wheels via a propeller shaft, final gear, or the like, and is always connected to the third ring gear R3.

  The first rotating member M1 is a rotating member that always connects the first ring gear R1 and the second ring gear R2 without interposing a friction element.

  The second rotating member M2 is a rotating member that always connects the second carrier PC2 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 third carrier PC3.

  The second clutch C2 is a second friction element that selectively connects the first carrier PC1 and the third carrier PC3.

  The third clutch C3 is a third friction element that selectively connects the first carrier PC1 and the second carrier PC2.

  The fourth clutch C4 is a fourth friction element that selectively connects the second sun gear S3 and the second carrier PC2.

  The first brake B1 is a fifth friction element that can lock the rotation of the second sun gear S2 with respect to the transmission case TC.

  The second brake B2 is a sixth friction element that can lock the rotation of the third carrier PC3 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 vertically 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 gear (1st) is achieved by simultaneously engaging the second clutch C2, the fourth clutch C4, and the second brake B2. As shown in FIG. 3, since the first planetary gear PG1 and the third planetary gear PG3 are involved in meshing, the total number of gear engagements at the first speed gear stage is 4 (= 2). Times + 2 times).

  The second speed (2nd) is achieved by simultaneously engaging the second clutch C2, 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 (= 2 times + 3 times + 2 times).

  As shown in FIG. 2, the third speed (3rd) is established 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 third planetary gear PG3 are involved in meshing, the total number of gear engagements at the third speed gear stage is 4 (= 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, since the first planetary gear PG1 and the second planetary gear PG2 are involved in meshing, the total number of gear engagements at the fourth speed gear stage is 5 (= 2). Times + 3 times).

  As shown in FIG. 2, the fifth speed (5th) is achieved by simultaneously engaging the first clutch C1, the second clutch C2, and the first brake B1. As shown in FIG. 3, since the second planetary gear PG2 and the third planetary gear PG3 are involved in meshing, the total number of gear engagements at the fifth speed gear stage is 5 (= 3). Times + 2 times).

  As shown in FIG. 2, the sixth gear (6th) is achieved by simultaneously engaging the first clutch C1, the second clutch C2, and the third clutch C3. 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 first clutch C1, the third clutch C3, and the first brake B1. As shown in FIG. 3, the number of gear meshes in the seventh speed gear stage is the total number of times because the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3 are involved in meshing. Is 7 times (= 2 times + 3 times + 2 times).

  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 is set to 2 because only the third planetary gear PG3 is involved in meshing.

  As shown in FIG. 2, the reverse speed (Rev) is achieved by simultaneously engaging the third clutch C3, 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 second clutch C2, the fourth clutch C4, and the second brake B2 are simultaneously engaged as shown by hatching in FIG.

  By simultaneously engaging the second clutch C2 and the second brake B2, the first carrier PC1 and the third carrier PC3 are fixed to the transmission case TC. With the engagement of the fourth clutch C4, the three rotating elements S2, PC2, R2 of the second planetary gear PG2 rotate integrally.

  Therefore, when the input rotational speed is input to the first sun gear S1 after passing through the input shaft IN, the first ring gear R1 is reversely rotated and decelerated in the first planetary gear PG1 fixed to the carrier. The rotation of the first ring gear R1 passes through the first rotating member M1, the second planetary gear PG2 (integrated rotation), and the second rotating member M2, and is input to the third sun gear S3. For this reason, in the third planetary gear PG3 fixed to the carrier, the rotation (reverse rotation deceleration) input to the third sun gear S3 is reversed, and the rotation is reduced to the normal rotation and output from the third ring gear R3. This output rotational speed (decelerated rotational speed lower than the input rotational speed) is transmitted as it is from the third ring gear R3 to the output shaft OUT, and the first speed gear stage is achieved.

(2nd gear)
At the second speed (2nd), the second clutch C2, the first brake B1, and the second brake B2 are simultaneously engaged as shown by hatching in FIG.

  By engaging the first brake B1, the second sun gear S2 is fixed to the transmission case TC. By simultaneous engagement of the second clutch C2 and the second brake B2, the first carrier PC1 and the third carrier PC3 are fixed to the transmission case TC.

  Therefore, when the input rotational speed is input to the first sun gear S1 after passing through the input shaft IN, the first ring gear R1 is reversely rotated and decelerated in the first planetary gear PG1 fixed to the carrier. The rotation of the first ring gear R1 passes through the first rotating member M1 and is input to the second ring gear R2. Then, in the second planetary gear PG2 fixed to the sun gear, the rotation input to the second ring gear R2 is reversely increased to be the rotation of the second carrier PC2. The rotation of the second carrier PC2 passes through the second rotating member M2 and is directly input to the third ring gear R3. For this reason, in the third planetary gear PG3 fixed to the carrier, the rotation (reverse rotation deceleration) input to the third sun gear S3 is reversed, and the rotation is reduced to the normal rotation and output from the third ring gear R3. This output rotational speed (lower rotational speed than the input rotational speed but higher than the first speed) is transmitted as it is from the third ring gear R3 to the output shaft OUT, and the second speed gear stage is achieved.

(3rd speed)
At the third speed (3rd), the second clutch C2, the fourth clutch C4, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By engaging the second clutch C2, the first carrier PC1 and the third carrier PC3 are directly connected. By simultaneously engaging the fourth clutch C4 and the first brake B1, the first ring gear R1 and the second planetary gear PG2 that are integrally connected via the first rotating member M1 and the second rotating member M2 are connected. Three rotating elements S2, PC2, R2 and a third sun gear S3 are fixed to the transmission case TC.

  Therefore, when the input rotational speed is input to the first sun gear S1 after passing through the input shaft IN, the first carrier PC1 is decelerated in the normal direction in the first planetary gear PG1 fixed to the ring gear. The rotation of the first carrier PC1 passes through the second clutch C2 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 (forward rotation deceleration) input to the third carrier PC3 is increased and output from the third ring gear R3. This output rotational speed (lower rotational speed than the input rotational speed but higher than the second speed) is transmitted as it is from the third ring gear R3 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.

  By simultaneously engaging the second clutch C2 and the third clutch C3, the first carrier PC1, the second carrier PC2, and the three rotating elements S3, PC3, and R3 of the third planetary gear PG3 are integrated. Rotate. As the first brake B1 is engaged, the second sun gear S2 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input to the first sun gear S1 after passing through the input shaft IN, the rotational speeds of the first carrier PC1 and the first ring gear R1 in the first planetary gear PG1 of the sun gear input are The sun planetary fixed second planetary gear PG2 rotates while being constrained by the rotational speeds of the second carrier PC2 and the second ring gear R2. The constraint condition at this time is that the first ring gear R1 and the second ring gear R2 maintain the same rotational speed via the first rotating member M1, and the first carrier PC1 and the first carrier PC1 via the third clutch C3. The condition is that the second carrier PC2 maintains the same rotational speed. Due to this rotational restraint relationship, the rotational speeds of the first carrier PC1 and the second carrier PC2 become the rotational speeds due to normal rotational deceleration with respect to the input rotational speed of the first sun gear S1. For this reason, the three rotating elements S3 of the third planetary gear PG3, which are directly connected to both the carriers PC1 and PC2, at the same rotational speed as the rotation (forward rotation deceleration) of the first carrier PC1 and the second carrier PC2. PC3 and R3 rotate together. For this reason, in the integrally rotating third planetary gear PG3, the output rotational speed (lower rotational speed than the input rotational speed but higher than the third speed) is directly transmitted from the third ring gear R3 to the output shaft OUT, and the fourth A fast gear is achieved.

(5th speed)
At the fifth speed (5th), the first clutch C1, the second clutch C2, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By simultaneously engaging the first clutch C1 and the second clutch C2, the three rotating elements S1, PC1, R1 of the first planetary gear PG1, the second ring gear R2, and the third carrier PC3 are integrated. Rotate. As the first brake B1 is engaged, the second sun gear S2 is fixed to the transmission case TC.

  Therefore, when the input rotation speed is input to the first sun gear S1 through the input shaft IN, the first planetary gear PG1 rotates integrally with the input rotation speed and passes through the first rotation member M1, The input rotational speed is input to the second ring gear R2 of the second planetary gear PG2 fixed to the sun gear. Then, in the second planetary gear PG2 fixed to the sun gear, the input rotational speed from the second ring gear R2 is increased and output from the second carrier PC2. The rotation of the second carrier PC2 passes through the second rotating member M2 and is directly input to the third sun gear S3. On the other hand, the input rotational speed is input to the third carrier PC3 after passing the first clutch C1. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotational speed input to the third sun gear S3 and the rotational speed input to the third carrier PC3 (= input rotational speed) are defined. Thus, the output rotational speed from the third ring gear R3 is determined. This output rotational speed (lower rotational speed than the input rotational speed but higher than the fourth speed) is transmitted as it is from the third ring gear R3 to the output shaft OUT, and the fifth speed gear stage is achieved.

(Sixth speed gear)
At the sixth speed (6th), the first clutch C1, the second clutch C2, and the third clutch C3 are simultaneously engaged as shown by hatching in FIG.

  By simultaneously engaging the first clutch C1, the second clutch C2, and the third clutch C3, the three rotating elements S1, PC1, R1 of the first planetary gear PG2 and the three rotating elements S2 of the second planetary gear PG2 are used. , PC2, R2 and the three rotating elements S3, PC3, R3 of the third planetary gear PG3 rotate together.

  Therefore, when the input rotational speed is inputted to the first sun gear S1 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 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 first clutch C1, the third clutch C3, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By engaging the first clutch C1, the first sun gear S1 and the third carrier PC3 are directly connected. By engaging the third clutch C3, the first carrier PC1 and the second carrier PC2 are directly connected. As the first brake B1 is engaged, the second sun gear S2 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input to the first sun gear S1 after passing through the input shaft IN, the rotational speeds of the first carrier PC1 and the first ring gear R1 in the first planetary gear PG1 of the sun gear input are The sun planetary fixed second planetary gear PG2 rotates while being constrained by the rotational speeds of the second carrier PC2 and the second ring gear R2. The constraint condition at this time is that the first ring gear R1 and the second ring gear R2 maintain the same rotational speed via the first rotating member M1, and the first carrier PC1 and the first carrier PC1 via the third clutch C3. The condition is that the second carrier PC2 maintains the same rotational speed. Due to this rotational restraint relationship, the rotational speeds of the first carrier PC1 and the second carrier PC2 become the rotational speeds due to normal rotational deceleration with respect to the input rotational speed of the first sun gear S1. The rotation of the second carrier PC2 passes through the second rotating member M2 and is directly input to the third sun gear S3. On the other hand, the input rotational speed is input to the third carrier PC3 after passing the first clutch C1. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotational speed input to the third sun gear S3 and the rotational speed input to the third carrier PC3 (= input rotational speed) are defined. Thus, the output rotational speed from the third ring gear R3 is determined. This output rotational speed (accelerated rotational speed higher than the input rotational speed) is directly transmitted from the third ring gear R3 to the output shaft OUT, and the seventh speed gear stage 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.

  By engaging the first clutch C1, the first sun gear S1 and the third carrier PC3 are directly connected. By simultaneously engaging the fourth clutch C4 and the first brake B1, the first ring gear R1 and the second planetary gear PG2 that are integrally connected via the first rotating member M1 and the second rotating member M2 are connected. Three rotating elements S2, PC2, R2 and a third sun gear S3 are fixed to the transmission case TC.

  Accordingly, the first planetary gear PG1 and the second planetary gear PG2 are not involved in achieving the eighth speed. For this reason, when the input rotational speed is inputted to the third carrier PC3 after passing through the input shaft IN and the first clutch C1, the third ring gear R3 is increased in the forward rotation in the third planetary gear PG3 fixed to the sun gear. Output at high speed. 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 third ring gear R3 to the output shaft OUT, and the eighth speed gear stage is achieved.

(Reverse speed shift stage)
At the reverse speed (Rev), 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 second carrier PC2 are directly connected. As the first brake B1 is engaged, the second sun gear S2 is fixed to the transmission case TC. As the second brake B2 is engaged, the third carrier PC3 is fixed to the transmission case TC.

  Therefore, when the input rotational speed is input to the first sun gear S1 after passing through the input shaft IN, the rotational speeds of the first carrier PC1 and the first ring gear R1 in the first planetary gear PG1 of the sun gear input are The sun planetary fixed second planetary gear PG2 rotates while being constrained by the rotational speeds of the second carrier PC2 and the second ring gear R2. The constraint condition at this time is that the first ring gear R1 and the second ring gear R2 maintain the same rotational speed via the first rotating member M1, and the first carrier PC1 and the first carrier PC1 via the third clutch C3. The condition is that the second carrier PC2 maintains the same rotational speed. Due to this rotational restraint relationship, the rotational speeds of the first carrier PC1 and the second carrier PC2 become the rotational speeds due to normal rotational deceleration with respect to the input rotational speed of the first sun gear S1. The rotation of the second carrier PC2 passes through the second rotating member M2 and is directly input to the third sun gear S3. For this reason, in the third planetary gear PG3 fixed to the carrier, the rotation input to the third sun gear S3 is reversed and output from the third ring gear R3. This output rotation speed (reverse rotation) is transmitted as it is from the third ring gear R3 to the output shaft OUT, and a reverse speed gear stage is achieved.

[Advantages by comparison with conventional technology]
FIG. 13 is a skeleton diagram showing a conventional automatic transmission. FIG. 14 is a diagram showing a fastening operation table for achieving 8 forward speeds and 2 reverse speeds by combining two of the six friction elements in the conventional automatic transmission. 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 automatic transmission of the conventional example (FIGS. 13 and 14), it can be said that the performance is equivalent with respect to the points listed below. .
(Basic configuration and speed change performance)
Both the automatic transmission of the conventional example and the automatic transmission of the first embodiment achieve eight forward speeds and one reverse speed with three planets and six friction elements.
(Transmission control performance)
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.

  However, the automatic transmission according to the first embodiment has advantages over the conventional automatic transmission in the points listed below.

(a) When selecting a planetary gear to be used for an automatic transmission for three planetary gears, there are single pinion planetary gears and double pinion planetary gears as options. Selection of a pinion planetary gear is preferred.

As shown in FIG. 13, the conventional automatic transmission uses a double pinion planetary gear and a Ravigneaux type planetary gear unit (one double pinion planetary gear and one single pinion planetary gear). In other words, because it uses two double pinion planetary gears,
-Gear transmission efficiency and gear noise are poor because the number of gear engagements increases.
・ Since the pinion gear diameter is reduced, durability reliability is reduced.
・ The number of parts increases, which increases costs.
There is a problem.

  On the other hand, in the case of the automatic transmission of the first embodiment, only the second planetary gear PG2 is a planetary gear by a double pinion with respect to the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3. The remaining two use planetary gears with a single pinion. For this reason, compared with the prior art which uses two double pinion planetary gears, the following items are advantageous.

(Reduction of gear meshing frequency)
In the case of the automatic transmission according to the first embodiment, the number of gear meshes is reduced as compared with the conventional example, the transmission efficiency of the gear is improved, and the gear noise is reduced.
That is, one set of double-pinion planetary gears has a meshing number of 3, whereas one set of single-pinion planetary gears has a meshing number of 2 because there is no meshing between the pinions. Therefore, in the case of Example 1 using a pair of double pinion planetary gears, the average meshing number is 4.25 as shown in FIG. In contrast, in the case of the conventional example using two sets of double pinion planetary gears, the average meshing number is 4.8 as shown in FIG. As a result, in the case of Example 1, the average number of meshes is reduced by 0.55 compared to the conventional example.

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

(Reduced number of parts)
In the case of the automatic transmission of the first embodiment, the number of parts is reduced, which is advantageous in terms of cost.
For example, in the case of a double pinion planetary gear, when four pairs 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 planetary gear, it is sufficient to arrange four pinions around the sun gear, and the number of parts is reduced by four. As a result, cost reduction can be achieved.

(b) Friction loss at each gear stage When friction elements are fastened to obtain each gear stage, friction loss cannot be avoided due to drag, etc. generated by the idling 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 dragging and the like by the four friction elements that idle is increased, 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 shift speeds, as shown in FIG. 2, three friction elements are simultaneously engaged at each shift speed. . For this reason, for example, the number of friction elements that idle at the first speed stage is three, such as the first clutch C1, the third clutch C3, and the first brake B1, at each speed stage. For this reason, compared with the conventional example, the friction loss at the friction element that idles can be suppressed to be small, and the transmission efficiency of the drive energy 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 range 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, as shown in FIG. 14, the value of RC = 6.711 (= 4.597 / 0.685). 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 = 0.585, the gear ratio of the second planetary gear PG2 is ρ2 = −0.360, When the gear ratio of the third planetary gear PG3 is set to ρ3 = 0.354, RC = 6.537 (= 4.829 / 0.739) is obtained while maintaining an appropriate gear ratio between adjacent gears.

  That is, while maintaining an appropriate gear ratio, the RC value can be made equal to that of the conventional example, and both the start performance at the lowest gear ratio and the high speed fuel consumption at the highest gear ratio are achieved. be able to. 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) Jumping shift For example, as a driving scene on a highway, etc., while the gear stage is traveling at a constant speed with the seventh speed on the overdrive side, an accelerator depression operation is performed with the intention of overtaking the vehicle ahead of the host vehicle. Then, a two-step jump down shift from the eighth speed to the fifth speed and a step-down shift by the two-stage jump down shift from the seventh speed to the fourth speed are performed.

In the case of the automatic transmission of the conventional example, both 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 are performed by releasing the two friction elements. This is a double-replacement shift in which two friction elements are fastened. Therefore, for example, 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 speed → sixth speed → fifth speed → fourth 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 maintained 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) About reverse power performance The first gear ratio and the reverse gear ratio are values that determine the start acceleration performance and the climbing performance. For example, if the ratio between the first gear ratio and the reverse gear ratio is not near 1, There is a difference in driving force when switching between decimals. 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, Rev1 / 1st = 0.882, Rev2 / 1st = 0.473, and in the case of Rev1 / 1st, the level that can prevent the driving force shortage at the time of reverse is What can be maintained, whether the first reverse speed (Rev1) or the second reverse speed (Rev2) is selected, the ratio of the first gear ratio to the reverse gear ratio is lower than 1, and is driven when switching between forward and reverse. There will be a power difference and the backward startability will be poor.

  On the other hand, in the case of the automatic transmission of the first embodiment, as shown in FIG. 2, Rev / 1st = 0.946, and the ratio between the first gear ratio and the reverse gear ratio is in the vicinity of one. For this reason, a driving force difference does not occur at the time of forward / reverse switching, and the backward startability is not deteriorated. That is, the vehicle can be operated without impairing the start acceleration performance and the climbing performance.

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 carrier PC1 supporting a first pinion P1 meshing with the first sun gear S1, and a first ring gear R1 meshing with the first pinion P1. The first planetary gear PG1, the second sun gear S2, the second ring gear R2, and the second double pinions P2s and P2r that mesh with the second sun gear S2 and the second ring gear R2, respectively. A second planetary gear PG2 composed of two carriers PC2, a third sun gear S3, a third carrier PC3 supporting a third pinion P3 meshing with the third sun gear S3, and the third pinion A third planetary gear PG3 comprising a third ring gear R3 meshing with P3, and six friction elements, and by appropriately fastening and releasing the six friction elements, the gear shifts to at least a forward 8-speed stage. Output torque from the input shaft IN to the output shaft OUT. In the possible automatic transmission, the input shaft IN is always connected to the first sun gear S1, the output shaft OUT is always connected to the third ring gear R3, and the first ring gear is connected. R1 and the second ring gear R2 are always connected to form a first rotating member M1, and the second carrier PC2 and the third sun gear S3 are always connected to form a second rotating member. M6 is configured, and the six friction elements include a first friction element (first clutch C1) that selectively connects the first sun gear S1 and the third carrier PC3; A second friction element (second clutch C2) for selectively connecting between the first carrier PC1 and the third carrier PC3, and a selection between the first carrier PC1 and the second carrier PC2. The third friction element (third clutch C3) to be connected to each other and the second planetary gear PG3 A fourth friction element (fourth clutch C4) for selectively connecting the two rotation elements, and a fifth friction element (first brake B1) capable of locking the rotation of the second sun gear S2. And a sixth friction element (second brake B2) capable of locking the rotation of the third carrier PC3, and at least the forward movement 8 is achieved by combining three of the six friction elements at the same time. Achieve 1st speed and reverse.
For this reason, while achieving 8 forward speeds with 3 planetary 6 friction elements, it is possible to achieve advantages in terms of gear efficiency, gear noise, durability reliability, and cost, and improvement in drive energy transmission efficiency.

(2) Of the six friction elements, the second friction element (second clutch C2) and the fourth friction element (fourth clutch C4) are at least in the eighth forward speed by a combination of three simultaneous engagements. And a first speed achieved by simultaneous engagement of the sixth friction element (second brake B2), the second friction element (second clutch C2), and the fifth friction element (first brake B1). A second speed achieved by simultaneous engagement of the sixth friction element (second brake B2), the second friction element (second clutch C2), the fourth friction element (fourth clutch C4), 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 first speed achieved by simultaneous engagement of the first friction element (first clutch C1), the second friction element (second clutch C2), and the fifth friction element (first brake B1); A sixth speed achieved by simultaneous engagement of the second friction element (first clutch C1), the second friction element (second clutch C2), and the third friction element (third clutch C3), and the first The seventh speed achieved by simultaneous engagement of the friction element (first clutch C1), the third friction element (third clutch C3), 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 the single replacement by fastening one friction element and releasing one friction element, which is advantageous in that the shift control is simplified. In addition, the RC value can be set so as to reach a required value for achieving both the start performance at the lowest gear ratio and the high speed fuel consumption at the highest gear ratio while maintaining an appropriate gear ratio. In addition, 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 Shifting with good response can be achieved by changing the weight.

(3) Of the six friction elements, the first reverse speed achieved by the combination of three simultaneous engagements is the third friction element (third clutch C3) and the fifth friction element (first brake B1). And the sixth friction element (second brake B2).
Therefore, the reverse gear ratio evaluation value (= reverse gear ratio / 1st gear ratio) can be set to a value near 1 even if a gear ratio that achieves an appropriate RC value and interstage ratio is selected. As a result, it is possible to prevent a difference in driving force when switching between forward and backward travel, and it is possible to ensure reverse start acceleration and climbing performance.

The second embodiment is an example in which the fourth clutch C4 is a clutch that directly connects the second carrier PC2 and the second ring gear R2 by fastening.

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 fourth clutch C4 of the first embodiment is a clutch that directly connects the second sun gear S2 and the second carrier PC2 by fastening (see FIG. 1), whereas the fourth clutch C4 of the second embodiment is fastened by fastening. The clutch directly connects the second carrier PC2 and the second ring gear R2.

  That is, the fourth clutch C4 of the second embodiment is a friction element that selectively connects between the two rotating elements of the third planetary gear PG3, similarly to the fourth clutch C4 of the first embodiment. For this reason, the second planetary gear PG2 can be integrally rotated or fixed integrally by engaging the fourth clutch C4 (first speed, third speed, and eighth speed). Since other configurations, operations, and effects are the same as those in the first embodiment, illustration and description thereof are omitted.

The third embodiment is an example in which the fourth clutch C4 is a clutch that directly connects the second sun gear S2 and the second ring gear R2 by fastening.

First, the configuration will be described.
FIG. 17 is a skeleton diagram illustrating the automatic transmission according to the third embodiment. Hereinafter, the planetary gear configuration and the friction element configuration of the automatic transmission according to the third embodiment will be described with reference to FIG.

  As shown in FIG. 17, the automatic transmission according to the third 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 fourth clutch C4 of the first embodiment is a clutch that directly connects the second sun gear S2 and the second carrier PC2 by engagement (see FIG. 1), and the fourth clutch C4 of the second embodiment is the second carrier by engagement. Whereas the clutch that directly connects PC2 and the second ring gear R2 (see FIG. 16), the fourth clutch C4 of the third embodiment is a clutch that directly connects the second sun gear S2 and the second ring gear R2 by fastening. did.

  That is, the fourth clutch C4 of the third embodiment is a friction element that selectively couples the two rotating elements of the third planetary gear PG3, similarly to the fourth clutch C4 of the first and second embodiments. For this reason, the third planetary gear PG3 can be rotated integrally by engaging the fourth clutch C4 (first speed, third speed, and eighth speed). Since other configurations, operations, and effects are the same as those in the first embodiment, illustration and description thereof are omitted.

  As mentioned above, although the automatic transmission of this invention has been demonstrated based on Example 1-3, 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 = 0.585, the gear ratio of the second planetary gear PG2 is ρ2 = −0.360, and the gear ratio of the third planetary gear PG3 is ρ3 = 0.354. An example was given. 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 value is not limited to the first embodiment.

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

PG1 first planetary gear
PG2 Second planetary gear
PG3 3rd planetary gear
IN input shaft
OUT output shaft
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 carrier that supports a first pinion that meshes with the first sun gear, and a first ring gear that meshes with the first pinion;
A second planetary gear comprising a second sun gear, a second ring gear, and a second carrier that supports a second double pinion that meshes with the second sun gear and the second ring gear, respectively.
A third planetary gear comprising a third sun gear, a third carrier supporting a third pinion meshing with the third sun gear, and a third ring gear meshing with the third pinion;
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 sun gear,
The output shaft is always connected to the third ring gear,
The first ring gear and the second ring gear are always connected to form a first rotating member,
The second carrier 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 third carrier;
A second friction element that selectively couples between the first carrier and the third carrier;
A third friction element for selectively connecting between the first carrier and the second rotating member;
A fourth friction element that selectively connects between two rotating elements of the second planetary gear;
A fifth friction element capable of locking the rotation of the second sun gear;
A sixth friction element capable of locking the rotation of the third carrier;
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 second friction element, the fourth friction element, and the sixth friction element;
A second speed achieved by simultaneous engagement of the second friction element, the fifth friction element and the sixth friction element;
A third speed achieved by simultaneous engagement of the second friction element, the fourth 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 first friction element, the second friction element, and the fifth friction element;
A sixth speed achieved by simultaneous engagement of the first friction element, the second friction element, and the third friction element;
A seventh speed achieved by simultaneous engagement of the first friction element, the third 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: In 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 third friction element, the fifth friction element, and the sixth friction element. Automatic transmission featured.

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