JP5214668B2 - Automatic transmission - Google Patents

Description

  The present invention relates to a stepped automatic transmission that is applied as a transmission of a vehicle.

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

Japanese Patent Laid-Open No. 2001-182785

However, in the conventional automatic transmission, two friction elements are fastened to achieve each of the eight forward speeds. For this reason, there are four friction elements that rotate idly in each shift stage, and there is a problem that friction loss at the idling friction elements is large, leading to deterioration in drive energy transmission efficiency.
In other words, 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 paying attention to the above-mentioned problem, and improves the transmission efficiency of driving energy by suppressing the friction loss generated at each gear stage while achieving the forward 8 speed or more with the three planetary 6 friction elements. An object of the present invention is to provide an automatic transmission capable of achieving the above.

In order to achieve the above object, an automatic transmission according to the present invention includes:
A first planetary gear comprising a first sun gear, a first ring gear, and a first carrier that supports the first sun gear and the first double pinion meshing with the first ring gear;
A second planetary gear comprising a second sun gear, a second ring gear, and a second carrier that supports the second sun gear and a second double pinion that meshes with the second ring gear;
A third planetary gear comprising a third sun gear, a third ring gear, and a third carrier that supports the third sun gear and a third double 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 output shaft is always connected to the third ring gear,
The first carrier is always fixed and constitutes a first fixed member,
The first sun gear and the second carrier are always connected to form a first rotating member,
The second sun gear and the third sun gear are always fastened to constitute a second rotating member,
The six friction elements are:
A first friction element that selectively connects between the first ring gear and the second rotating member;
A second friction element that selectively couples between the third carrier and the first rotating member;
A third friction element that selectively couples between the second ring gear and the third carrier;
A fourth friction element that selectively couples between the input shaft and the third carrier;
A fifth friction element that selectively connects between the input shaft and the first ring gear;
A sixth friction element that selectively locks rotation of the second ring gear;
Composed of
Of the six friction elements, at least eight forward speeds and one reverse speed are achieved by a combination of three simultaneous fastenings.

  Therefore, in the automatic transmission of the present invention, at least eight forward speeds and one reverse speed are achieved by combining three simultaneous engagements among the six friction elements. For this reason, at each shift stage, there are three friction elements that idle, and the friction loss at the friction element that idles can be kept small. As a result, the transmission energy transmission efficiency can be improved by suppressing the friction loss that occurs at each shift stage while achieving the forward 8th speed or higher with the three planetary six friction elements.

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 9 forward speeds and 1 reverse speed by the combination of three simultaneous fastening among the six friction elements in the automatic transmission of Example 1. It is a figure which shows the maximum torque sharing ratio table | surface of each friction element in the automatic transmission of Example 1. FIG. 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. 10 is an explanatory diagram showing a shift operation at a ninth speed (9th) 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 maximum torque sharing ratio table | surface of each friction element in the automatic transmission of a prior art example.

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

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

  As shown in FIG. 1, the automatic transmission according to the first embodiment includes a first planetary gear PG1, a second planetary gear PG2, a third planetary gear PG3, an input shaft IN, an output shaft OUT, A first rotating member M1, a second rotating member M2, a first fixed member F1, a first clutch C1 (first friction element), a second clutch C2 (second friction element), Third clutch C3 (third friction element), fourth clutch C4 (fourth friction element), fifth clutch C5 (fifth friction element), and first brake B1 (sixth friction element) And a transmission case TC.

  The first planetary gear PG1 is a double pinion type planetary gear having first double pinions P1s and P1r, and meshes with the first sun gear S1, the pinion P1s meshed with the first sun gear S1, and the pinion P1s. The first carrier PC1 that supports the pinion P1r and the first ring gear R1 that meshes with the pinion P1r.

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

  The third planetary gear PG3 is a double pinion type planetary gear having third double pinions P3s and P3r. The third planetary gear PG3 meshes with the third sun gear S3, the pinion P3s meshed with the third sun gear S3, and the pinion P3s. It consists of a third carrier PC3 that supports the pinion P3r and a third ring gear R3 that meshes with the pinion P3r.

  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.

  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 sun gear S1 and the second carrier PC2 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 fixing member F1 is a member that always fixes the first carrier PC1 to the transmission case TC.

  The first clutch C1 is a first friction element that selectively connects the first ring gear R1 and the second rotating member M2.

  The second clutch C2 is a second friction element that selectively connects the third carrier PC3 and the first rotating member M1.

  The third clutch C3 is a third friction element that selectively connects the second ring gear R2 and the third carrier PC3.

  The fourth clutch C4 is a fourth friction element that selectively connects the input shaft IN and the third carrier PC3.

  The fifth clutch C5 is a fifth friction element that selectively connects the input shaft IN and the first ring gear R1.

  The first brake B1 is a sixth friction element that can lock the rotation of the second ring gear R2 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 illustrating a fastening operation table for achieving 9 forward speeds and 1 reverse speed by combining three simultaneous fastenings among six friction elements in the automatic transmission according to the first embodiment. Hereinafter, based on FIG. 2, a description will be given of a shift configuration that establishes each shift stage of the automatic transmission according to the first embodiment.

  The automatic transmission according to the first embodiment has 9 forward speeds and 1 reverse speed as described below by combining three of the six friction elements C1, C2, C3, C4, C5, and B1 simultaneously. To achieve.

  As shown in FIG. 2, the first speed (1st) is achieved by simultaneously engaging the fourth clutch C4, the fifth clutch C5, and the first brake B1.

  As shown in FIG. 2, the second speed (2nd) is achieved by simultaneously engaging the second clutch C2, the fourth clutch C4, and the first brake B1.

  As shown in FIG. 2, the third speed (3rd) is achieved by simultaneously engaging the first clutch C1, the fourth clutch C4, and the first brake B1.

  As shown in FIG. 2, the fourth speed (4th) is achieved by simultaneously engaging the first clutch C1, the second clutch C2, and the fourth clutch C4.

  As shown in FIG. 2, the fifth speed (5th) is achieved by simultaneously engaging the first clutch C1, the third clutch C3, and the fourth clutch C4.

  As shown in FIG. 2, the sixth speed (6th) is achieved by simultaneously engaging the first clutch C1, the fourth clutch C4, and the fifth clutch C5.

  The seventh speed (7th) is achieved by simultaneously engaging the first clutch C1, the third clutch C3, and the fifth clutch C5, as shown in FIG.

  As shown in FIG. 2, the eighth speed (8th) is achieved by simultaneous engagement of the first clutch C1, the second clutch C2, and the fifth clutch C5.

  As shown in FIG. 2, the ninth speed (9th) is achieved by simultaneously engaging the second clutch C2, the third clutch C3, and the fifth clutch C5.

  As shown in FIG. 2, the reverse speed (Rev) is achieved by simultaneously engaging the third clutch C3, the fifth clutch C5, and the first brake B1.

  FIG. 3 is a diagram illustrating a maximum torque sharing ratio table of each friction element in the automatic transmission according to the first embodiment. Here, the “torque sharing ratio” is a ratio of torque acting on each friction element expressed as a ratio to the input torque when the input torque is “1”. The “maximum torque sharing ratio” indicates the maximum value among the torque sharing ratios of the friction elements at the respective speed stages including the first speed to the ninth speed and the reverse speed. As the maximum torque sharing ratio increases, the torque acting on the friction element increases, so the number of friction plates increases and the size increases. Hereinafter, the maximum torque sharing ratio of each friction element of the automatic transmission according to the first embodiment will be described with reference to FIG.

  The first clutch C1 is engaged when the third speed, the fourth speed, the fifth speed, the sixth speed, the seventh speed, and the eighth speed are achieved, but the torque sharing ratio is maximized. Is the seventh speed, and the maximum torque sharing ratio at that time is 0.709.

  The second clutch C2 is engaged when the second speed, fourth speed, eighth speed, and ninth speed are achieved. The torque sharing ratio is the largest at the ninth speed. Yes, the maximum torque sharing ratio at that time is 0.748.

  The third clutch C3 is engaged when the fifth speed, the seventh speed, the ninth speed, and the reverse speed are achieved. The torque sharing ratio is the largest at the reverse speed. The maximum torque sharing ratio at that time is 1.709.

  The fourth clutch C4 is engaged when the first speed, the second speed, the third speed, the fourth speed, the fifth speed, and the sixth speed are achieved, but the torque sharing ratio becomes the largest. Is the first speed gear stage, and the maximum torque sharing ratio at that time is 2.410.

  The fifth clutch C5 is engaged when the first, sixth, seventh, eighth, ninth, and reverse speeds are achieved, but the torque sharing ratio is the largest. The first speed gear stage, and the maximum torque sharing ratio at that time is 1.410.

  The first brake B1 is engaged when the first speed, the second speed, the third speed, and the reverse speed are achieved. The torque sharing ratio is the largest at the reverse speed. The maximum torque sharing ratio at that time is 3.512.

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 fourth clutch C4, the fifth clutch C5, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By engaging the fourth clutch C4, the input shaft IN and the third carrier PC3 are directly connected. By engaging the fifth clutch C5, the input shaft IN and the first ring gear R1 are directly connected. As the first brake B1 is engaged, the second ring gear R2 is fixed to the transmission case TC.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation speed is input to the first ring gear R1 and the third carrier PC3. Therefore, in the first planetary gear PG1 fixed to the carrier, the input rotation speed is increased and output from the first sun gear S1. The rotation of the first sun gear S1 is input as it is to the second carrier PC2 via the first rotating member M1. Therefore, in the second planetary gear PG2 fixed to the ring gear, the rotation of the second carrier PC2 is reversed and output from the second sun gear S2. The rotation of the second sun gear S2 is inputted as it is to the third sun gear S3 via the second rotating member M2. For this reason, in the third planetary gear PG3 having two inputs and one output, the number of rotations of the third sun gear S3 and the number of rotations of the third carrier PC3 (= input number of rotations) are defined. The rotation speed of R3 is determined. The output rotation speed (= deceleration rotation speed lower than the input rotation speed) from the third ring gear R3 is directly transmitted 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 fourth clutch C4, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  The input shaft IN, the first sun gear S1, the second carrier PC2, and the third carrier PC3 are directly connected by the simultaneous engagement of the second clutch C2 and the fourth clutch C4 and the first rotating member M1. As the first brake B1 is engaged, the second ring gear R2 is fixed to the transmission case TC.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation speed is input to the first sun gear S1, the second carrier PC2, and the third carrier PC3. For this reason, in the second planetary gear PG2 fixed to the ring gear, the input rotation is decelerated and output from the second sun gear S2. The rotation of the second sun gear S2 is inputted as it is to the third sun gear S3 via the second rotating member M2. For this reason, in the third planetary gear PG3 having two inputs and one output, the number of rotations of the third sun gear S3 and the number of rotations of the third carrier PC3 (= input number of rotations) are defined. The rotation speed of R3 is determined. The output rotation speed (= deceleration rotation speed lower than the input rotation speed and higher than the first speed) from the third ring gear R3 is directly transmitted 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 fourth clutch C4, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By engaging the fourth clutch C4, the input shaft IN and the third carrier PC3 are directly connected. With the simultaneous engagement of the first clutch C1 and the first brake B1, and the first and second rotating members M1, M2 and the first fixed member F1, the three rotating elements S1, PC1, R1 of the first planetary gear PG1 The three rotating elements S2, PC2, and R2 of the second planetary gear PG2 are integrally fixed to the transmission case TC, and the third sun gear S3 is fixed to the transmission case TC.

  Accordingly, when the input rotational speed is input to the third carrier PC3 after passing through the input shaft IN, the input rotation is decelerated and output from the third ring gear R3 in the third planetary gear PG3 fixed to the sun gear. The output rotation speed (= deceleration rotation speed lower than the input rotation speed and higher than the second speed) from the third ring gear R3 is directly transmitted to the output shaft OUT, and the third speed gear stage is achieved.

(4th gear)
At the fourth speed (4th), the first clutch C1, the second clutch C2, and the fourth clutch C4 are simultaneously engaged as shown by hatching in FIG.

  The input shaft IN, the first sun gear S1, the second carrier PC2, and the third carrier PC3 are directly connected by the simultaneous engagement of the second clutch C2 and the fourth clutch C4 and the first rotating member M1. The first ring gear R1, 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.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation speed is input to the first sun gear S1, the second carrier PC2, and the third carrier PC3. Therefore, in the first planetary gear PG1 fixed to the carrier, the input rotation is decelerated and output from the first ring gear R1. The rotation of the first ring gear R1 is directly input to the third sun gear S3 via the first clutch C1 and the second rotating member M2. For this reason, in the third planetary gear PG3 having two inputs and one output, the number of rotations of the third sun gear S3 and the number of rotations of the third carrier PC3 (= input number of rotations) are defined. The rotation speed of R3 is determined. The output rotation speed (= deceleration rotation speed lower than the input rotation speed and higher than the third speed) from the third ring gear R3 is directly transmitted to the output shaft OUT, and the fourth speed gear stage is achieved.

(5th speed)
At the fifth speed (5th), the first clutch C1, the third clutch C3, and the fourth clutch C4 are simultaneously engaged as shown by hatching in FIG.

  By simultaneously engaging the third clutch C3 and the fourth clutch C4, the input shaft IN, the second ring gear R2, and the third carrier PC3 are directly connected. The first ring gear R1, the second sun gear S2, and the third sun gear S3 are directly connected by the engagement of the first clutch C1 and the second rotating member M2.

  Therefore, when the input shaft IN rotates at the input rotational speed, the input rotational speed is input to the second ring gear R2 and the third carrier PC3, and at the same time, the second sun gear S2 and the second second gear S2 of the second planetary gear PG2. The carrier PC2 rotates while being restricted by the rotation of the first sun gear S1 and the first ring gear R1 of the first planetary gear PG1 fixed to the carrier. The constraint condition at this time is that the first sun gear S1 and the second carrier PC2 maintain the same rotational speed via the first rotating member M1, and the first ring gear R1 and the first ring gear R1 via the first clutch C1. The condition is that the second sun gear S2 maintains the same rotational speed. Due to this rotational constraint relationship, the rotational speed of the second sun gear S2 becomes the rotational speed obtained by decelerating the input rotation. The rotation of the second sun gear S2 is inputted as it is to the third sun gear S3 via the second rotating member M2. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotation speed of the third sun gear S3 and the rotation speed of the third carrier PC3 (= input rotation speed) are defined, whereby the third ring gear. The rotation speed of R3 is determined. The output rotation speed (= deceleration rotation speed lower than the input rotation speed and higher than the fourth speed) from the third ring gear R3 is directly transmitted 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 fourth clutch C4, and the fifth clutch C5 are simultaneously engaged as shown by hatching in FIG.

  By the simultaneous engagement of the first clutch C1, the fourth clutch C4, and the fifth clutch C5 and the second rotating member M2, two rotating elements S3 and PC3 are directly connected to the third planetary gear PG3 to form the third planetary gear. The three rotating elements S3, PC3, R3 of PG3 are brought into a state of rotating integrally, and the input shaft IN, the first ring gear R1, the second sun gear S2, and the third planetary gear PG3 are directly connected.

  Therefore, when the input shaft IN rotates at the input rotation speed, the third planetary gear PG3 rotates integrally at the input rotation speed. For this reason, the output rotational speed from the third ring gear R3 (= the same rotational speed as the input rotational speed from the input shaft IN) is directly transmitted to the output shaft OUT, and the sixth speed gear stage (direct connection) with a gear ratio of 1 is established. Shift stage) is achieved.

(7th speed)
At the seventh speed (7th), the first clutch C1, the third clutch C3, and the fifth clutch C5 are simultaneously engaged as shown by hatching in FIG.

  The input shaft IN, the first ring gear R1, the second sun gear S2, and the third sun gear S3 are directly connected by the simultaneous engagement of the first clutch C1 and the fifth clutch C5 and the second rotating member M2. By engaging the third clutch C3, the second ring gear R2 and the third carrier PC3 are directly connected.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation speed is input to the first ring gear R1, the second sun gear S2, and the third sun gear S3. Therefore, in the first planetary gear PG1 fixed to the carrier, the input rotation speed is increased and output from the first sun gear S1. The rotation of the first sun gear S1 is input as it is to the second carrier PC2 via the first rotating member M1. For this reason, in the second planetary gear PG2 having two inputs and one output, the rotation speed of the second sun gear S2 (= input rotation speed) and the rotation speed of the second carrier PC2 are defined, whereby the second ring gear. The rotation speed of R2 is determined. The rotation of the second ring gear R2 is directly input to the third carrier PC3 via the third clutch C3. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotation speed of the third sun gear S3 (= input rotation speed) and the rotation speed of the third carrier PC3 are defined, whereby the third ring gear. The rotation speed of R3 is determined. The output rotation speed from the third ring gear R3 (= acceleration speed higher than the input rotation speed) is transmitted as it is 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 second clutch C2, and the fifth clutch C5 are simultaneously engaged as shown by hatching in FIG.

  The input shaft IN, the first ring gear R1, the second sun gear S2, and the third sun gear S3 are directly connected by the simultaneous engagement of the first clutch C1 and the fifth clutch C5 and the second rotating member M2. The first sun gear S1, the second carrier PC2, and the third carrier PC3 are directly connected by the engagement of the second clutch C2 and the first rotating member M1.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation speed is input to the first ring gear R1, the second sun gear S2, and the third sun gear S3. Therefore, in the first planetary gear PG1 fixed to the carrier, the input rotation speed is increased and output from the first sun gear S1. The rotation of the first sun gear S1 is directly input to the third carrier PC3 via the first rotating member M1 and the second clutch C2. For this reason, in the third planetary gear PG3 having two inputs and one output, the rotation speed of the third sun gear S3 (= input rotation speed) and the rotation speed of the third carrier PC3 are defined, whereby the third ring gear. The rotation speed of R3 is determined. The output rotation speed from the third ring gear R3 (= the input rotation speed and the speed increase rotation speed higher than the seventh speed) is directly transmitted to the output shaft OUT, and the eighth speed gear stage is achieved.

(9th speed)
At the ninth speed (9th), the second clutch C2, the third clutch C3, and the fifth clutch C5 are simultaneously engaged as shown by hatching in FIG.

  By engaging the fifth clutch C5, the input shaft IN and the first ring gear R1 are directly connected. By the simultaneous engagement of the second clutch C2 and the third clutch C3 and the first and second rotating members M1, M2, the two rotating elements PC2, R2 are directly connected in the second planetary gear PG2, and the second planetary gear PG2 The three rotating elements S2, PC2, R2 are brought into a state of rotating integrally, and the two rotating elements S3, PC3 are directly connected in the third planetary gear PG3, so that the three rotating elements S3, PC3 of the third planetary gear PG3 are connected. , R3 are rotated together, and the first sun gear S1, the second planetary gear PG2, and the third planetary gear PG3 are directly connected.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation speed is input to the first ring gear R1. Therefore, in the first planetary gear PG1 fixed to the carrier, the input rotation speed is increased and output from the first sun gear S1. The rotation of the first sun gear S1 is directly input to the third planetary gear PG3 via the second clutch C2, the third clutch C3, and the first and second rotating members M1 and M2, and the third planetary gear PG3. Rotates together with the rotation speed of the first sun gear S1. For this reason, the output rotational speed (= the input rotational speed and the speed increasing rotational speed higher than the eighth speed) of the third planetary gear PG3 from the third ring gear R3 is directly transmitted to the output shaft OUT, and the ninth speed A shift stage is achieved.

(Reverse speed shift stage)
At the reverse speed (Rev), the third clutch C3, the fifth clutch C5, and the first brake B1 are simultaneously engaged as shown by hatching in FIG.

  By engaging the fifth clutch C5, the input shaft IN and the first ring gear R1 are directly connected. By simultaneously engaging the third clutch C3 and the first brake B1, the second ring gear R2 and the third carrier PC3 are fixed to the transmission case TC.

  Therefore, when the input shaft IN rotates at the input rotation speed, the input rotation speed is input to the first ring gear R1. Therefore, in the first planetary gear PG1 fixed to the carrier, the input rotation speed is increased and output from the first sun gear S1. The rotation of the first sun gear S1 is input as it is to the second carrier PC2 via the first rotating member M1. For this reason, in the second planetary gear PG2 fixed to the ring gear, rotation opposite to the rotation direction of the second carrier PC2 is output from the second sun gear S2. The rotation of the second sun gear S2 is inputted as it is to the third sun gear S3 via the second rotating member M2. Therefore, in the third planetary gear PG3 fixed to the carrier, the rotation of the third sun gear S3 is increased and output from the third ring gear R3. For this reason, the output rotational speed from the third ring gear R3 (= the decelerating rotational speed opposite to the input rotational speed and lower than the input rotational speed) is directly transmitted to the output shaft OUT, and the reverse speed gear stage is achieved. The

[Advantages by comparison with conventional technology]
FIG. 14 is a skeleton diagram showing a conventional automatic transmission. FIG. 15 is a diagram showing a fastening operation table for achieving the 8th forward speed and the 2nd reverse speed by combining two of the six friction elements in the conventional automatic transmission. FIG. 16 is a diagram showing a maximum torque sharing ratio table of each friction element 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 to 3) and the automatic transmission of the conventional example (FIGS. 14 to 16), it can be said that the performance is equivalent with respect to the points listed below. .
(Basic configuration)
Both the automatic transmission of the conventional example and the automatic transmission of the first embodiment are configured by three planets and six friction elements.
(Transmission control performance)
In the automatic transmission of the conventional example and the automatic transmission of the first embodiment, the shift to the adjacent shift stage and the shift to the one-step jump shift stage are called release of one friction element and engagement of one friction element. This is achieved by a single overhanging shift.
(Tooth ratio)
The conventional automatic transmission and the automatic transmission of the first embodiment are both absolute values of the gear ratios ρ1, ρ2, and ρ3 of the first planetary gear PG1, the second planetary gear PG2, and the third planetary gear PG3. Is set within the range of 0.3 to 0.65.

  On the other hand, `` (a) Friction loss at each gear stage '', `` (b) transmission performance '', `` (c) gear ratio width '', `` (d) reverse power performance '' and `` (e) unit layout '' listed below The automatic transmission according to the first embodiment has an advantage over the conventional automatic transmission.

(a) When a friction loss friction element at each shift stage is engaged to obtain each shift stage, friction loss cannot be avoided due to oil dragging or the like caused by a slipping friction element (release element). However, as an automatic transmission Is preferably as the friction loss is smaller.

  In the case of the automatic transmission of the conventional example, 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, there are four friction elements that idle at each speed, such as the second clutch C2, the third clutch C3, the fourth clutch C4, and the first brake B1, that idle at the first speed. It becomes. 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. That is, for example, when the conventional automatic transmission is applied to an engine vehicle, the friction loss due to the four friction elements that idle causes a deterioration in fuel efficiency.

  On the other hand, in the case of the automatic transmission of the first embodiment, in order to achieve each of the nine forward speeds, as shown in FIG. 2, three friction elements are simultaneously engaged at each 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 second clutch C2, and the third clutch C3. 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. That is, for example, when the automatic transmission according to the first embodiment is applied to an engine vehicle, fuel efficiency can be improved.

(b) Speed change performance In the case of the automatic transmission of the conventional example, the eight-speed forward speed is achieved by three planets and six friction elements, whereas in the case of the automatic transmission of the first embodiment, three planets and six The 9th forward speed is achieved by the friction element.

  For this reason, since the choice of a gear ratio spreads and the driving force according to the condition of a vehicle can be output, a fuel consumption can be improved. Further, the increase in the number of gears reduces the interval between the gears, so that the driving force step between gears and the gear shift shock can be suppressed.

(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 automatic transmission of the conventional example, as shown in FIG. 15, the tooth number ratio of the double pinion type planetary gear is ρ1 = −0.375, and the tooth number ratio of the Ravioni type planetary gear unit is ρ2 = 0.500, ρ3 = −0.375. In this case, RC = 6.397 (= 4.267 / 0.667). On the other hand, in the automatic transmission according to the first embodiment, as shown in FIG. 2, the gear ratio of the first planetary gear PG1 is ρ1 = −0.617, and the gear ratio of the second planetary gear PG2 is ρ2 = −. When the gear ratio of the third planetary gear PG3 is 0.658 and ρ3 = −0.410, RC = 6.623 (= 4.084 / 0.617) 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 higher than that of the conventional example, and both start performance at the lowest gear ratio and high speed fuel consumption at the highest gear ratio can be ensured. . Here, “appropriate gear ratio” means that the gear ratio at each gear stage is plotted, and when the plotted points are connected by a line, the characteristic is smooth from the low gear side toward the high gear side. It means that a characteristic line can be drawn that changes in a flat state after being lowered by a gradient.

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

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

  In the case of the conventional automatic transmission, as shown in FIG. 15, Rev1 / 1st = 0.750 and Rev2 / 1st = 0.469. Therefore, in the case of Rev1 / 1st, that is, if the reverse 1st speed (Rev1) is selected, the level that can prevent the driving force from being insufficient during reverse can be maintained, but if the 2nd reverse speed (Rev2) is selected, the 1st gear The ratio between the ratio and the reverse gear ratio is a value significantly lower than 1, and a driving force difference may occur when switching between forward and reverse travel, and reverse startability may be deteriorated.

  On the other hand, in the case of the automatic transmission of the first embodiment, as shown in FIG. 2, Rev / 1st = 0.709, and the ratio of the first gear ratio to the reverse gear ratio is the same as that of the first reverse gear of the conventional example. It becomes. For this reason, it is possible to prevent the driving force from being insufficient when switching between forward and backward travel, and the backward startability is not deteriorated. That is, the vehicle can be operated without impairing the start acceleration performance and the climbing performance.

(e) Unit layout In the automatic transmission of the conventional example, as shown in FIG. 16, the largest one of the maximum torque sharing ratios of the friction elements (the first clutch C1 to the second brake B2) is the second brake B2. 4.8000. On the other hand, the maximum torque sharing ratio of each friction element (the first clutch C1 to the first brake B1) of the automatic transmission according to the first embodiment, as shown in FIG. B1 is 3.512. For this reason, the number of friction plates in the friction element can be reduced, and manufacturing can be performed at a low cost. In addition, the size of each friction element (the first clutch C1 to the first brake B1) can be suppressed, and the unit layout Can be prevented.

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

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

(1) From the first sun gear S1, the first ring gear R1, and the first carrier PC1 that supports the first sun gear S1 and the first double pinions P1s and P1r meshing with the first ring gear R1. The first planetary gear PG1, the second sun gear S2, the second ring gear R2, and the second double pinions P2s and P2r meshing with the second sun gear S2 and the second ring gear R2 are supported. A second planetary gear PG2 comprising two carriers PC2, 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 planetary gear PG3 comprising a third carrier PC3, and six friction elements, and by appropriately fastening and releasing the six friction elements, at least a forward 8-speed gear stage is achieved. Change the speed and output torque from the input shaft IN to the output shaft OUT. In the possible automatic transmission, the output shaft OUT is always connected to the third ring gear R3, and the first carrier PC1 is always fixed to form the first fixed member F1, The first sun gear S1 and the second carrier PC2 are always connected to form a first rotating member M1, and the second sun gear S2 and the third sun gear S3 are always fastened. The second rotating member is configured as M2, and the six friction elements are a first friction element (first clutch) that selectively connects the first ring gear R1 and the second rotating member M2. C1), a second friction element (second clutch C2) that selectively connects the third carrier PC3 and the first rotating member M1, the second ring gear R2, and the third A third friction element (third clutch C3) for selectively connecting the carrier PC3, A fourth friction element (fourth clutch C4) that selectively connects between the shaft IN and the third carrier PC3, and a first friction gear that selectively connects between the input shaft IN and the first ring gear R1. 5 friction elements (fifth clutch C5) and a sixth friction element (first brake B1) capable of locking the rotation of the second ring gear R2, and among the six friction elements, By combining three simultaneous fastenings, at least 8 forward speeds and 1 reverse speed were achieved.
For this reason, it is possible to improve the transmission efficiency of the drive energy by suppressing the friction loss generated at each shift stage while achieving the forward 8th speed or higher with the three planetary 6 friction elements.

(2) Among the six friction elements, the ninth forward speed by the combination of three simultaneous engagements is the fourth friction element (fourth clutch C4), the fifth friction element (fifth clutch C5), and the The first speed achieved by simultaneous engagement of the sixth friction element (first brake B1), the second friction element (second clutch C2), the fourth friction element (fourth clutch C4), and the second Second speed achieved by simultaneous engagement of six friction elements (first brake B1), the first friction element (first clutch C1), the fourth friction element (fourth clutch C4), and the sixth The third speed achieved by simultaneous engagement of the first friction element (first brake B1), the first friction element (first clutch C1), the second friction element (second clutch C2), and the fourth speed. Fourth speed achieved by simultaneous engagement of the friction element (fourth clutch C4), and the first friction element (First clutch C1), fifth speed achieved by simultaneous engagement of the third friction element (third clutch C3) and the fourth friction element (fourth clutch C4), and the first friction element ( The sixth speed achieved by simultaneous engagement of the first clutch C1), the fourth friction element (fourth clutch C4) and the fifth friction element (fifth clutch C5), and the first friction element (first 7th speed achieved by simultaneous engagement of the first clutch C1), the third friction element (third clutch C3) and the fifth friction element (fifth clutch C5), and the first friction element (first The eighth speed achieved by simultaneous engagement of the clutch C1), the second friction element (second clutch C2) and the fifth friction element (fifth clutch C5); and the second friction element (second clutch). C2), the third friction element (third clutch C3), and the fifth friction element (fifth clutch C5). A ninth speed that is established by simultaneous engagements of was made of configurations.
For this reason, the choice of gear ratio spreads, the driving force according to the condition of a vehicle can be output, and a fuel consumption can be improved. Further, the increase in the number of gears reduces the interval between the gears, so that the driving force step between gears and the gear shift shock can be suppressed. 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.

(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 (fifth clutch C5). The sixth friction element (first brake B1) is achieved by simultaneous engagement.
Therefore, the reverse gear ratio evaluation value (= reverse gear ratio / 1st gear ratio) can be set to a value near 1 even if a gear ratio that achieves an appropriate RC value and interstage ratio is selected. As a result, it is possible to prevent the occurrence of a driving force difference when switching between forward and backward travel, and it is possible to ensure reverse start acceleration and climbing performance.

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

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

  In the first embodiment, an example of an automatic transmission that is applied to an FR engine vehicle in which the input / output shaft is coaxially arranged is shown. The present invention can also be applied as an automatic transmission for various vehicles such as cars. Moreover, it is also useful as a transmission of a vehicle equipped with a diesel engine having a narrower engine speed range as a power source than a gasoline engine and having a lower torque when compared with the same displacement.

PG1 first planetary gear
S1 First sun gear
PC1 first carrier
R1 1st ring gear
PG2 Second planetary gear
S2 Second sun gear
PC2 Second carrier
R2 Second ring gear
PG3 3rd planetary gear
S3 3rd sun gear
PC3 Third carrier
R3 3rd ring gear
IN input shaft
OUT output shaft
F1 First fixed member
M1 First rotating member
M2 Second rotating member
C1 First clutch (first friction element)
C2 Second clutch (second friction element)
C3 3rd clutch (3rd friction element)
C4 4th clutch (4th friction element)
C5 5th clutch (5th friction element)
B1 First brake (sixth friction element)
TC transmission case

Claims (3)

A first planetary gear comprising a first sun gear, a first ring gear, and a first carrier that supports the first sun gear and the first double pinion meshing with the first ring gear;
A second planetary gear comprising a second sun gear, a second ring gear, and a second carrier that supports the second sun gear and a second double pinion that meshes with the second ring gear;
A third planetary gear comprising a third sun gear, a third ring gear, and a third carrier that supports the third sun gear and a third double 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 output shaft is always connected to the third ring gear,
The first carrier is always fixed and constitutes a first fixed member,
The first sun gear and the second carrier are always connected to form a first rotating member,
The second sun gear and the third sun gear are always fastened to constitute a second rotating member,
The six friction elements are:
A first friction element that selectively connects between the first ring gear and the second rotating member;
A second friction element that selectively couples between the third carrier and the first rotating member;
A third friction element that selectively couples between the second ring gear and the third carrier;
A fourth friction element that selectively couples between the input shaft and the third carrier;
A fifth friction element that selectively connects between the input shaft and the first ring gear;
A sixth friction element that selectively locks rotation of the second ring gear;
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, the forward 9-speed by the combination of three simultaneous fastenings is:
A first speed achieved by simultaneous engagement of the fourth friction element, the fifth friction element, and the sixth friction element;
A second speed achieved by simultaneous engagement of the second friction element, the fourth friction element, and the sixth friction element;
A third speed achieved by simultaneous engagement of the first friction element, the fourth friction element, and the sixth friction element;
A fourth speed achieved by simultaneous engagement of the first friction element, the second friction element, and the fourth friction element;
A fifth speed achieved by simultaneous engagement of the first friction element, the third friction element, and the fourth friction element;
A sixth speed achieved by simultaneous engagement of the first friction element, the fourth friction element, and the fifth 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 second friction element, and the fifth friction element;
A ninth speed achieved by simultaneous engagement of the second friction element, the third 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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