Shift mechanism of stepped automatic transmission for automobiles

Abstract

To provide a speed change mechanism for a stepped automatic transmission for an automobile, which does not use a clutch and thereby enables a simpler structure, a smaller size, a lighter weight, and a lower cost. [Solution] A gear train is formed by connecting a plurality of planetary gears (PG11, PG12, PG21, PG0). Each of the planetary gears includes a sun gear (S11, S12, S21, S1, S2), a ring gear (R11, R12, R21, R0), and a carrier (C11, C12, C21, C0) that rotatably supports the plurality of pinion gears. Each of the planetary gears is operated only by a brake (B11, B12, B21, B01, B02) without using a clutch.

Description

[Technical field]

[0001] The present invention relates to a transmission mechanism for a stepped automatic transmission for an automobile. [Background technology]

[0002] A known speed change mechanism for a stepped automatic transmission for automobiles and the like is equipped with a plurality of double pinion planetary gear sets, and is provided with a clutch interposed between rotating elements for engaging the rotating elements together, and a brake interposed between the rotating elements and a case for stopping the rotating elements, as friction elements for controlling engagement and release during gear shifting (see, for example, Patent Document 1). A transmission using an incomplete planetary gear mechanism having two components, one central gear and one planet carrier, has also been proposed (see, for example, Patent Document 2). [Prior art documents] [Patent documents]

[0003] [Patent Document 1] JP 2007-321770 A [Patent Document 2] Special Publication No. 2022-527927 Summary of the Invention [Problem to be solved by the invention]

[0004] A wet multi-plate clutch is generally used as the above clutch. However, the clutch mechanism has the following problems: the rotating clutch drum is heavy, a mechanism for canceling residual pressure when released is required, the clutch support mechanism is complex and heavy, and space-saving, the oil passage structure for supplying pressure oil to the rotating body is complex, a seal ring is required to prevent pressure oil leakage, and the degree of freedom in layout of the entire automatic transmission is reduced.

[0005] SUMMARY OF THE PRESENT EMBODIMENTS The present invention has an object to provide a speed change mechanism for a stepped automatic transmission for automobiles which does not use a clutch and thereby enables a simplified structure, a smaller size, a lighter weight, and a reduced cost. [Means for solving the problem]

[0006] In order to achieve the above object, the present invention provides a speed change mechanism for a stepped automatic transmission for an automobile, in which a gear train is formed by connecting (linking) a plurality of planetary gears, each of which includes a sun gear, a ring gear, and a carrier that rotatably supports a plurality of pinion gears, and each of the plurality of planetary gears is operated only by a brake without using a clutch. Here, the speed change mechanism of the present invention is intended for a stepped automatic transmission for an automobile using an engine as a power source. The speed change mechanism of the present invention includes a plurality of planetary gears, but each planetary gear is a general one, and is highly reliable with no problems in processing and assembly. The plurality of planetary gears are linked together to function as a multi-stage transmission by selecting and combining brakes to be applied. In addition, each planetary gear can be arranged almost in close contact with each other, and does not require a space as in the case of using a pinion shaft gear (Patent Document 2) that has a plurality of gears and is integrated with them. The control function structures of existing stepped automatic transmissions for automobiles are either clutch only or a combination of clutch and brake, and there is no brake only structure. The inventor believes that the present invention is unique in the structure of stepped automatic transmissions for automobiles.

[0007] Here, it is preferable that the plurality of planetary gears include a compound planetary gear, and a plurality of shift speeds are input to one input member of the compound planetary gear. More specifically, a first output member of a first planetary gear and a second output member of a second planetary gear are connected to one input member of the compound planetary gear. Then, the input from the first planetary gear and the input from the second planetary gear are switched so that either one is input to the one input member by operating a first brake provided on the first planetary gear or a second brake provided on the second planetary gear.

[0008] It is also preferable to have a plurality of reverse gears.

[0009] It is also preferable that the plurality of planetary gears include two single-row type double pinion planetary gears connected in series, the power flow of the reverse gear is formed by the two single-row type double pinion planetary gears, and the input rotation direction and output rotation direction are the same when the two single-row type double pinion planetary gears are viewed as a group.

[0010] It is also preferable that the vehicle further includes a clutch hub sleeve, and that the clutch hub sleeve is used as a reverse gear selection system.

[0011] In addition, it is preferable that the plurality of planetary gears include two planetary gear groups each consisting of a plurality of single-row planetary gears and each having a reverse rotation output function, and that the two planetary gear groups are connected in series to provide a forward rotation output as a transmission.

[0012] In addition, it is preferable that the plurality of planetary gears include a first planetary gear group consisting of a plurality of single-row planetary gears and having a reverse rotation output function, and a second planetary gear group consisting of another plurality of single-row planetary gears and having a forward rotation output function, and that the first planetary gear group and the second planetary gear group are connected in series to provide a reverse rotation output as a transmission.

[0013] It is also preferable that the plurality of planetary gears include a plurality of single-row planetary gears, and that the brake controlling each of the single-row planetary gears brakes the ring gear of each of the single-row planetary gears for all of the plurality of single-row planetary gears. Effect of the Invention

[0014] According to the gear shift mechanism of the stepped automatic transmission for automobiles described in claim 1, each of the planetary gears is operated only by brakes without using clutches, which simplifies the structure, makes it compact, lightweight, and reduces costs, and also improves the freedom of layout of the entire automatic transmission.

[0015] According to the speed change mechanism of a stepped automatic transmission for automobiles described in claim 2, a plurality of shift speeds are input to one input member of the compound planetary gear, so that the gear ratio can be increased with a simple structure.

[0016] The Ravigneaux planetary gear, which is a type of compound planetary gear, is a compact compound planetary gear that combines a single pinion planetary gear and a double pinion planetary gear. For example, in the case of forward movement, if the small sun gear S1 is the input member, two reduction ratios can be obtained when the large sun gear S2 is fixed and when the carrier is fixed. Here, if multiple speed change speeds (for example, two) are input to the small sun gear S1, it is possible to obtain four reduction ratios, or 2 x 2.

[0017] According to the shift mechanism of the stepped automatic transmission for automobiles described in claim 3, there are multiple reverse gears. Generally, there is only one reverse gear, but in this case the gear ratio is equal to that of the first forward gear, and the driving force is large. This poses an issue of safety when backing into parking or entering a garage, but if there are multiple reverse gears, it is possible to use a gear with appropriate driving force even when reversing. In the present invention, since a clutch is not used and only a brake is used, the entire structure of the automatic transmission is compact and there is a large degree of freedom in layout, so that multiple reverse gears can be easily realized in terms of layout.

[0018] According to the transmission mechanism of a stepped automatic transmission for automobiles described in claim 4, the power flow of the reverse gear is formed by two single-row double pinion planetary gears connected in series. In each single-row double pinion planetary gear, when one of the sun gear and the carrier is the input and the other is the output, the input rotation direction and the output rotation direction are opposite to each other. Therefore, when the two single-row double pinion planetary gears are viewed as a group, the output rotation is obtained in the same direction as the input rotation direction.

[0019] According to the speed change mechanism of a stepped automatic transmission for automobiles described in claim 5, a clutch hub sleeve is used as a reverse gear selection method, so that a reverse gear can be realized by a simple mechanism.

[0020] According to the gear change mechanism of the stepped automatic transmission for automobiles described in claim 6, two planetary gear groups each having a reverse output function and each consisting of a plurality of single row planetary gears are connected in series to provide a forward output as a transmission, whereby, for example, when traveling forward, an output rotation in the same direction as the input rotation direction can be obtained.

[0021] According to the gear change mechanism of the stepped automatic transmission for automobiles described in claim 7, a first planetary gear group consisting of a plurality of single row planetary gears and having a reverse rotation output function and a second planetary gear group consisting of another plurality of single row planetary gears and having a forward rotation output function are connected in series to provide a reverse output as a transmission, whereby, for example, when reversing, an output rotation in the opposite direction to the input rotation direction is obtained.

[0022] According to the transmission mechanism of a stepped automatic transmission for automobiles described in claim 8, the brake brakes the ring gear of each of the single-row planetary gears for all of the single-row planetary gears. Generally, brakes are of the wet multi-plate type, with rotating discs (the side to be braked) and fixed discs packed alternately, and the fixed discs are fitted into the case. Since the ring gear is located on the outside close to the case, a configuration for braking the ring gear is advantageous in terms of layout. [Brief description of the drawings]

[0023] [Figure 1] FIG. 2 is an explanatory diagram showing the configuration of a single-row single-pinion planetary gear. [Diagram 2] FIG. 2 is an explanatory diagram showing the configuration of a single-row double pinion planetary gear. [Diagram 3] FIG. 2 is an explanatory diagram showing the configuration of a Ravigneaux type planetary gear. [Figure 4] 4 is a table showing output reduction ratios corresponding to various operating modes of the Ravigneaux type planetary gear; [Figure 5A] The left figure is a simplified explanatory diagram of a wet-type multi-plate brake, and the right figure is a corresponding skeleton diagram. [Figure 5B] 1 is an explanatory diagram of a clutch hub sleeve, the right side showing the untightened state and the left side showing the tightened state. [Figure 6] FIG. 2 is a diagram showing how to represent a single-row single pinion planetary gear, a single-row double pinion planetary gear, a Ravigneaux type planetary gear, and a wet-type multi-plate brake in a skeleton diagram. [Figure 7A]1 is a skeleton diagram of a transmission mechanism FB0605 for a stepped automatic transmission for an automobile according to a first embodiment (hereinafter, also referred to simply as a "transmission mechanism"); [Figure 7B] 2 is a table showing brake application and gear ratios at each speed change stage of the speed change mechanism FB0605 according to the first embodiment (hereinafter also simply referred to as the "brake application table"). [Figure 8A] FIG. 13 is a skeleton diagram of a speed change mechanism FA0603 according to a second embodiment. [Figure 8B] 13 is a brake application table for the speed change mechanism FA0603 according to the second embodiment. [Figure 9A] FIG. 11 is a skeleton diagram of a speed change mechanism RB0803 according to a third embodiment. [Figure 9B] 13 is a brake application table for the speed change mechanism RB0803 according to the third embodiment. [Figure 10A] FIG. 13 is a skeleton diagram of a speed change mechanism FA0501 according to a fourth embodiment. [Figure 10B] 13 is a brake application table for the speed change mechanism FA0501 according to the fourth embodiment. [Figure 11] 4 is a skeleton diagram and a brake application table of a speed change mechanism FA0601 according to another embodiment. [Figure 12] 4 is a skeleton diagram and a brake application table of a speed change mechanism FA0602 according to another embodiment. [Figure 13] 4 is a skeleton diagram and a brake application table of a speed change mechanism FA0604 according to another embodiment. [Figure 14] 4 is a skeleton diagram and a brake application table of a speed change mechanism FA0605 according to another embodiment. [Figure 15] 4 is a skeleton diagram and a brake application table of a speed change mechanism FA0606 according to another embodiment. [Figure 16] 4 is a skeleton diagram and a brake application table of a speed change mechanism FA0801 according to another embodiment. [Figure 17] 13 is a skeleton diagram and a brake application table of a speed change mechanism FA0802 according to another embodiment. [Figure 18] 13 is a skeleton diagram and a brake application table of a speed change mechanism FA0803 according to another embodiment. [Figure 19]13 is a skeleton diagram and a brake application table of a speed change mechanism FA0804 according to another embodiment. [Figure 20] 13 is a skeleton diagram and a brake application table of a speed change mechanism FA0805 according to another embodiment. [Figure 21] 13 is a skeleton diagram and a brake application table of a speed change mechanism FA0806 according to another embodiment. [Figure 22] 13 is a skeleton diagram and a brake application table of a speed change mechanism FA0901 according to another embodiment. [Figure 23] 13 is a skeleton diagram and a brake application table of a speed change mechanism FA0902 according to another embodiment. [Figure 24] 13 is a skeleton diagram and a brake application table of a speed change mechanism FA0903 according to another embodiment. [Diagram 25] 13 is a skeleton diagram and a brake application table of a transmission mechanism FB0601 according to another embodiment. [Figure 26] 13 is a skeleton diagram and a brake application table of a transmission mechanism FB0602 according to another embodiment. [Figure 27] 13 is a skeleton diagram and a brake application table of a transmission mechanism FB0603 according to another embodiment. [Figure 28] 13 is a skeleton diagram and a brake application table of a transmission mechanism FB0604 according to another embodiment. [Figure 29] 13 is a skeleton diagram and a brake application table of a transmission mechanism FB0801 according to another embodiment. [Diagram 30] 13 is a skeleton diagram and a brake application table of a transmission mechanism FB0802 according to another embodiment. [Diagram 31] 13 is a skeleton diagram and a brake application table of a transmission mechanism FB0901 according to another embodiment. [Diagram 32] 5 is a skeleton diagram and a brake application table of a transmission mechanism FB1001 according to another embodiment. [Diagram 33] 13 is a skeleton diagram and a brake application table of a speed change mechanism FC0601 according to another embodiment. [Diagram 34] 13 is a skeleton diagram and a brake application table of a speed change mechanism FC0602 according to another embodiment. [Diagram 35]13 is a skeleton diagram and a brake application table of a speed change mechanism FC0603 according to another embodiment. [Diagram 36] 13 is a skeleton diagram and a brake application table of a speed change mechanism FC0604 according to another embodiment. [Figure 37] 13 is a skeleton diagram and a brake application table of a speed change mechanism FC0801 according to another embodiment. [Figure 38] 13 is a skeleton diagram and a brake application table of a speed change mechanism FC0802 according to another embodiment. [Figure 39] 13 is a skeleton diagram and a brake application table of a speed change mechanism FC0901 according to another embodiment. [Diagram 40] 13 is a skeleton diagram and a brake application table of a speed change mechanism FC0902 according to another embodiment. [Diagram 41] 4 is a skeleton diagram and a brake application table of a transmission mechanism FD0601 according to another embodiment. [Diagram 42] 13 is a skeleton diagram and a brake application table of a transmission mechanism FD0801 according to another embodiment. [Diagram 43] 13 is a skeleton diagram and a brake application table of a transmission mechanism FD0901 according to another embodiment. [Diagram 44] 13 is a skeleton diagram and a brake application table of a transmission mechanism FD0902 according to another embodiment. [Diagram 45] 13 is a skeleton diagram and a brake application table of a transmission mechanism FD0903 according to another embodiment. [Diagram 46] 13 is a skeleton diagram and a brake application table of a speed change mechanism RA0801 according to another embodiment. [Figure 47] 13 is a skeleton diagram and a brake application table of a speed change mechanism RA0802 according to another embodiment. [Figure 48] 13 is a skeleton diagram and a brake application table of a speed change mechanism RA0803 according to another embodiment. [Figure 49] 13 is a skeleton diagram and a brake application table of a speed change mechanism RA0804 according to another embodiment. [Figure 50] 13 is a skeleton diagram and a brake application table of a speed change mechanism RA0805 according to another embodiment. [Figure 51]13 is a skeleton diagram and a brake application table of a speed change mechanism RA0901 according to another embodiment. [Figure 52] 13 is a skeleton diagram and a brake application table of a speed change mechanism RA0902 according to another embodiment. [Diagram 53] 13 is a skeleton diagram and a brake application table of a speed change mechanism RA0903 according to another embodiment. [Figure 54] 11 is a skeleton diagram and a brake application table of a speed change mechanism RA1001 according to another embodiment. [Figure 55] 13 is a skeleton diagram and a brake application table of a speed change mechanism RA1201 according to another embodiment. [Figure 56] 4 is a skeleton diagram and a brake application table of a speed change mechanism RB0601 according to another embodiment. [Figure 57] 4 is a skeleton diagram and a brake application table of a speed change mechanism RB0801 according to another embodiment. [Figure 58] 13 is a skeleton diagram and a brake application table of a speed change mechanism RB0802 according to another embodiment. [Figure 59] 13 is a skeleton diagram and a brake application table of a speed change mechanism RB0901 according to another embodiment. [Figure 60] 13 is a skeleton diagram and a brake application table of a speed change mechanism RB0902 according to another embodiment. [Figure 61] 13 is a skeleton diagram and a brake application table of a speed change mechanism RB0903 according to another embodiment. [Figure 62] 13 is a skeleton diagram and a brake application table of a speed change mechanism RB0904 according to another embodiment. [Figure 63] 13 is a skeleton diagram and a brake application table of a speed change mechanism RB0905 according to another embodiment. [Figure 64] 4 is a skeleton diagram and a brake application table of a speed change mechanism RB1001 according to another embodiment. [Figure 65] 10 is a skeleton diagram and a brake application table of a speed change mechanism RB1002 according to another embodiment. [Figure 66] 4 is a skeleton diagram and a brake application table of a speed change mechanism RB1003 according to another embodiment. [Figure 67]4 is a skeleton diagram and a brake application table of a speed change mechanism RB1201 according to another embodiment. [Figure 68] 12 is a skeleton diagram and a brake application table of a speed change mechanism RB1202 according to another embodiment. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0024] Before describing embodiments of the present invention, some general information will first be provided to facilitate understanding. <Speed ​​relationship equation for single-row single-pinion planetary gear> Generally, in the single-row single-pinion planetary gear shown in Figure 1, the sun gear tooth number is ZS, the ring gear tooth number is ZR, the sun gear rotation speed is ωS, the carrier rotation speed is ωC, and the ring gear rotation speed is ωR. Also, λ is the sun gear tooth number ratio (λ=ZS / ZR). The speed relation equation for a single-row single-pinion planetary gear is expressed by the following equation. Note that here, it is called "single-row type" in contrast to the compound planetary gear described later. (1+λ)ωC =λωS+ωR (Equation 1)

[0025] <Speed ​​relationship equation for single-row double pinion planetary gear> Generally, in the single-row double pinion planetary gear shown in Figure 2, the sun gear tooth number is ZS, the ring gear tooth number is ZR, the sun gear rotation speed is ωS, the carrier rotation speed is ωC, and the ring gear rotation speed is ωR. λ is the sun gear tooth number ratio (λ=ZS / ZR). The speed relation equation of the single-row double pinion planetary gear is expressed by the following equation. (1-λ)ωC = -λωS+ωR (Eq. 2)

[0026] <Speed ​​relationship equation for Ravigneaux type planetary gear> Generally, in the Ravigneaux type planetary gear shown in Figure 3, if the number of teeth of the sun gear S1 is ZS1, the rotational speed of the sun gear S1 is ωS1, the number of teeth of the sun gear S2 is ZS2, the rotational speed of the sun gear S2 is ωS2, the number of teeth of the ring gear is ZR, the rotational speed of the ring gear is ωR, the rotational speed of the carrier is ωC, the tooth ratio of the sun gear S1 is α = ZS1 / ZR, and the tooth ratio of the sun gear S2 is β = ZS2 / ZR, the following speed relationship holds for the Ravigneaux type planetary gear. Note that the inner pinion that meshes with the sun gear S1 is also called the short pinion, and the outer pinion that meshes with the ring gear is also called the long pinion. The Ravigneaux type planetary gear is an example of a compound type planetary gear. For column S1 (1-α)·ωC = -α·ωS1 + ωR (Equation 3A) For the S2 column, (1+β)·ωC = β·ωS2 + ωR (Eq. 3B)

[0027] Furthermore, if ωi is the input rotation speed of the AT (automatic transmission) main body via the joint, ωo is the output rotation speed of the AT main body (final pinion rotation speed), and i is the gear ratio as the AT (i = ωi / ωo), the output reduction ratio (= ωi / ωR) when the input to the Ravigneaux planetary gear is the reduced input and the ring gear is the output is given by each equation in the table of Figure 4. Here, the input reduction ratio to the sun gear S1 is J1, the input reduction ratio to the sun gear S2 is J2, and the input reduction ratio to the carrier is J3. Note that each equation in Figure 4 is a known equation.

[0028] <Brakes> The brake used in this embodiment will be described. The left diagram of FIG. 5A is a simplified explanatory diagram of the brake B, and the right diagram is a corresponding skeleton diagram. The brake B in this embodiment is a wet-type multi-plate type, in which a fixed side disk D1 (the side that applies the brake) and a rotating side disk D2 (the side that is braked) are alternately packed. The fixed side disk D1 is fitted into a case K. The rotating side disk D2 is fitted into a ring gear R. Friction material is attached to both sides of the rotating side disk D2. Although not shown, the fixed side disk D1 and the rotating side disk D2 are pressed against each other by a mechanism including a piston and a retainer (stopping plate) that press the disks together, thereby braking (stopping) the ring gear R. Note that FIG. 5A shows a single pinion planetary gear as an example, and only one side of the rotation axis X is shown for simplification. The ring gear R meshes with the pinion P, and the pinion P has a pinion shaft PS. The pinion shaft PS is supported by a carrier C. The pinion P meshes with the sun gear S.

[0029] <Clutch hub sleeve> A clutch hub sleeve (hereinafter, also referred to as CHS) mechanism will be described. CHS is generally applied to manual transmissions. FIG. 5B is a simplified cross-sectional view of CHS. As shown in FIG. 5B, CHS includes a hub H and a sleeve SL. Rotating body A such as a gear includes a meshing portion AK having a spline groove on its outer periphery. Hub H includes a meshing portion HK having a spline groove on its outer periphery. Sleeve SL includes a meshing portion SK having a spline groove on its inner periphery. Sleeve SL is movable in the axial direction relative to rotating body A and hub H. In the state shown in the right diagram of FIG. 5B, sleeve SL meshes only with hub H. In this state, rotating body A and hub H can rotate independently. In the state shown in the left diagram of FIG. 5B after the sleeve SL moves in the axial direction, sleeve SL meshes with both rotating body A and hub H. In this state, rotating body A and hub H rotate integrally in the same direction at the same rotation speed.

[0030] Fig. 6 shows the representation of a single-row single pinion planetary gear (SP-PG), a single-row double pinion planetary gear (DP-PG), a composite planetary gear (Ravigneaux type in this case), and a wet-type multi-plate brake, which are used in the skeleton diagrams of the present application. In the skeleton diagrams of the present application, to avoid complexity, the rotating bodies are basically represented on one side with the rotation axis X as the center.

[0031] First Embodiment A transmission mechanism FB0605 for a stepped automatic transmission for an automobile according to a first embodiment of the present invention will be described with reference to Figs. 7A and 7B. Fig. 7A is a skeleton diagram of a transmission mechanism FB0605 for a front-wheel drive 6-speed forward / 1-speed reverse automatic transmission. To avoid complexity, the rotating body is illustrated on one side (upper half) with the rotation axis X as the center.

[0032] The transmission mechanism FB0605 includes three double pinion planetary gears PG11, PG12, PG21 and a Ravigneaux planetary gear PG0. A main gear train including the double pinion planetary gears PG11, PG12, PG21 and the Ravigneaux planetary gear PG0 is arranged on one axis (rotation axis X). The double pinion planetary gears PG11, PG12, PG21 and the Ravigneaux planetary gear PG0 each have a known configuration. The double pinion planetary gear PG11 includes a sun gear S11, a ring gear R11, and a carrier C11. The double pinion planetary gear PG12 includes a sun gear S12, a ring gear R12, and a carrier C12. The carrier C11 and the carrier C12 are connected to each other. The double pinion planetary gear PG21 includes a sun gear S21, a ring gear R21, and a carrier C21. The Ravigneaux planetary gear PG0 includes two sun gears S1 and S2, a ring gear R0, and a carrier C0. The sun gear S1 is connected to the carrier C11. The sun gear S2 is connected to the carrier C21.

[0033] The tooth ratio of the double pinion planetary gear PG11 is λ11 = 36 / 90 = 0.400. Here, the number "36" displayed on the sun gear S11 in FIG. 7A indicates the number of teeth of the sun gear S11, and the number "90" displayed on the ring gear R11 indicates the number of teeth of the ring gear R11. The same is true for the numbers displayed on the other gears. The tooth ratio of the double pinion planetary gear PG12 is λ12 = 45 / 90 = 0.500. The tooth ratio of the double pinion planetary gear PG21 is λ21 = 45 / 90 = 0.500.

[0034] The transmission mechanism FB0605 further includes brakes B11, B12, B21, B01, and B02. Each of these brakes is a wet multi-plate type in which a rotating disc (the side to be braked) and a fixed disc are alternately packed, and the fixed disc is fitted into the case K. The brake B11 stops the rotation of the ring gear R11 of the double pinion planetary gear PG11. The brake B12 stops the rotation of the ring gear R12 of the double pinion planetary gear PG12. The brake B21 stops the rotation of the ring gear R21 of the double pinion planetary gear PG21. The brake B01 stops the rotation of the carrier C0 of the Ravigneaux planetary gear PG0. The brake B02 stops the rotation of the carrier C21 of the double pinion planetary gear PG21 and the sun gear S2 of the Ravigneaux planetary gear PG0.

[0035] The input coupling mechanism IN is a fluid coupling, a torque converter, a wet clutch, or the like. The input rotation from the input coupling mechanism IN is input to the sun gear S12 of the double pinion planetary gear PG12, the sun gear S11 of the double pinion planetary gear PG11, and the sun gear S21 of the double pinion planetary gear PG21. The sun gear S1 of the Ravigneaux planetary gear PG0 is connected to the carrier C11 of the double pinion planetary gear PG11 and the carrier C12 of the double pinion planetary gear PG12. That is, there are two inputs to the sun gear S1 of the Ravigneaux planetary gear PG0. The sun gear S2 of the Ravigneaux planetary gear PG0 is connected to the carrier C21 of the double pinion planetary gear PG21. That is, there is one input to the sun gear S2 of the Ravigneaux planetary gear PG0. In addition, there is no input to the carrier C0 of the Ravigneaux planetary gear PG0.

[0036] The transmission mechanism FB0605 is connected to the output section G0. The output section G0 includes a transmission mechanism output gear G1, an output idle gear G2, and an output gear set G3. The output gear set G3 drives a differential device (not shown). The differential device (not shown) includes a final reduction gear. In addition, T in the figure indicates the rearmost part of the gear train that determines the gear ratio (gear ratio). Hereinafter, it is also referred to as T section or T shaft. The gear ratio at T section is shown in the rightmost column of FIG. 7B (the same applies to FIGS. 8B, 9B, and 10B described later).

[0037] The rotation of ring gear R0 of Ravigneaux planetary gear PG0 is transmitted to output gear G1 of the speed change mechanism, and this rotation is transmitted in turn to output idle gear G2, output gear set G3, and a differential device (not shown). Also, in the brake application table of Figure 7B, for example, in the column for brake B01, there is a circle (O) in the column for the gear positions reverse (R), 1st gear (circled number 1), and 2nd gear (circled number 2), indicating that the brake is applied (braking state). On the other hand, in the column for brake B01, there is no circle (O) in the column for the gear positions 3rd gear (circled number 3) through 6th gear (circled number 6), indicating that the brake is not applied (non-braking state).

[0038] <When retreating> As shown in FIG. 7B, in reverse (R), the brakes B01 and B21 are engaged. Because the brakes B11 and B12 are not engaged, the double pinion planetary gears PG11 and PG12 are not activated, and the rotation from the input coupling mechanism IN is input to the sun gear S21 of the double pinion planetary gear PG21. Because the ring gear R21 is stopped, the carrier C21 rotates. The rotation of the carrier C21 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. Because the carrier C0 is stopped, the rotation of the sun gear S2 is transmitted to the ring gear R0. The rotation of the ring gear R0 is output to the output section G0.

[0039] Here, for the double pinion planetary gear PG21, the ring gear rotation speed ωR = 0, so from the above formula (2), the carrier rotation speed ωC = -1.000ωS and the reduction ratio J21 = -1.000, which is input to the sun gear S2 of the Ravigneaux planetary gear PG0. The carrier C0 of the Ravigneaux planetary gear PG0 is fixed (ωC = 0). Note that in the rotation speed and reduction ratio, a minus sign means that the rotation is reversed.

[0040] The reverse gear ratio iR is calculated as follows: Sun gear S1 rotation speed: ωS1=-(β / α)·(1 / J21)=-(0.333 / 0.267)·(1 / -1.000)ωi=1.247ωi Sun gear S2 rotation speed: ωS2=(1 / J21)ωi=(ωi / -1.000)= -1.000ωi Output speed (= ring gear R0 speed): ωo=ωR=-(β / J2)·ωi=(0.333 / -1.000) ωi=0.333ωi (see Figure 4(g)) Reverse gear ratio: iR=(ωi / ωo)= (ωi / 0.333ωi) ≒3.00

[0041] <1st gear> As shown in FIG. 7B, in first gear (number 1 in a circle), brakes B01 and B11 are engaged. Rotation from the input coupling mechanism IN enters the sun gear S11 of the double pinion planetary gear PG11, and because the ring gear R11 is stopped, the carrier C11 rotates. The rotation of the carrier C11 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Because the carrier C0 is stopped, the rotation of the sun gear S1 is transmitted to the ring gear R0. The rotation of the ring gear R0 is output to the output section G0.

[0042] Here, for double pinion planetary gear PG11, the ring gear rotation speed ωR = 0, so from the above formula (2), the carrier rotation speed ωC = -0.667ωS and the reduction ratio J11 = -1.500, which is input to the sun gear S1 of Ravigneaux planetary gear PG0. The carrier C0 of Ravigneaux planetary gear PG0 is fixed (ωC = 0).

[0043] The gear ratio i1 in 1st gear is calculated as follows: Sun gear S1 rotation speed: ωS1=(1 / J11)ωi= (ωi / -1.500) =-0.667ωi Sun gear S2 rotation speed: ωS2=-(α / β)·(1 / J11)ωi = -(0.267 / 0.333)·(1 / -1.500)ωi=0.535ωi Output speed (= ring gear R0 speed): ωo=ωR=(α / J11) ·ωi =(0.267 / -1.500) ωi=-0.178ωi (see Figure 4(a)) Gear ratio in 1st gear: i1=(ωi / ωo)= (ωi / -0.178ωi) ≒-5.62

[0044] <2nd gear> As shown in FIG. 7B, in second gear (number 2 in a circle), brakes B01 and B12 are engaged. Rotation from the input coupling mechanism IN enters the sun gear S12 of the double pinion planetary gear PG12, and because the ring gear R12 is stopped, the carrier C12 rotates. Because the carrier C12 and carrier C11 are connected, the rotation of the carrier C12 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Because the carrier C0 is stopped, the rotation of the sun gear S1 is transmitted to the ring gear R0. The rotation of the ring gear R0 is output to the output section G0.

[0045] Here, for the double pinion planetary gear PG12, the ring gear rotation speed ωR = 0, so from the above formula (2), the carrier rotation speed ωC = -1.000ωS and the reduction ratio J12 = -1.000, which is input to the sun gear S1 of the Ravigneaux planetary gear PG0. The carrier C0 of the Ravigneaux planetary gear PG0 is fixed (ωC = 0).

[0046] The gear ratio i2 in second gear is calculated as follows: Sun gear S1 rotation speed: ωS1=(1 / J12)ωi= (ωi / -1.000) =-1.000ωi Sun gear S2 rotation speed: ωS2=-(α / β)·(1 / J12)ωi =-(0.267 / 0.333)·(1 / -1.000)ωi=0.802ωi Output speed (= ring gear R0 speed): ωo=ωR=(α / J12)·ωi=(0.267 / -1.000) ωi=-0.267ωi (see Figure 4(a)) 2nd gear ratio: i2=(ωi / ωo)= (ωi / -0.267ωi) ≒-3.75

[0047] <3rd gear> As shown in FIG. 7B, in third gear (number 3 in a circle), brakes B02 and B11 are engaged. Rotation from the input coupling mechanism IN enters the sun gear S11 of the double pinion planetary gear PG11, and because the ring gear R11 is stopped, the carrier C11 rotates. The rotation of the carrier C11 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. The rotation of the sun gear S1 is transmitted from the short pinion to the long pinion, and because the sun gear S2 is stopped, the carrier C0 and ring gear R0 rotate. The rotation of the ring gear R0 is output to the output section G0.

[0048] Here, for double pinion planetary gear PG11, ring gear rotation speed ωR = 0, so from the above formula (2), carrier rotation speed ωC = -0.667ωS, and reduction ratio J11 = -1.500, which is input to sun gear S1 of Ravigneaux planetary gear PG0. Sun gear S2 of Ravigneaux planetary gear PG0 is fixed (ωS2 = 0).

[0049] The gear ratio i3 in 3rd gear is calculated as follows: Sun gear S1 rotation speed: ωS1=(1 / J11)ωi= (ωi / -1.500) =-0.667ωi Sun gear S2 rotation speed: ωS2=0 Carrier C0 rotation speed: ωC={α / (α+β)}·(1 / J11)ωi =(0.267 / 0.600)·(1 / -1.500)ωi=-0.297ωi Output speed (= ring gear R0 speed): ωo=ωR={α(1+β) / (α+β)}·(1 / J11)ωi =(0.356 / 0.600)·(1 / -1.500)ωi=-0.396ωi (See (b) in Figure 4) 3rd gear ratio: i3=(ωi / ωo)= (ωi / -0.396ωi) ≒-2.53

[0050] <4th gear> As shown in FIG. 7B, in fourth gear (number 4 in a circle), brakes B02 and B12 are engaged. Rotation from the input coupling mechanism IN enters the sun gear S12 of the double pinion planetary gear PG12, and because the ring gear R12 is stopped, the carrier C12 rotates. The rotation of the carrier C12 is input to the sun gear S1 of the Ravigneaux planetary gear PG0 via the carrier C11. The rotation of the sun gear S1 is transmitted from the short pinion to the long pinion, and because the sun gear S2 is stopped, the carrier C0 and ring gear R0 rotate. The rotation of the ring gear R0 is output to the output section G0.

[0051] Here, for double pinion planetary gear PG12, ring gear rotation speed ωR = 0, so from the above formula (2), carrier rotation speed ωC = -1.000ωS and reduction ratio J12 = -1.000, which is input to sun gear S1 of Ravigneaux planetary gear PG0. Sun gear S2 of Ravigneaux planetary gear PG0 is fixed (ωS2 = 0).

[0052] The gear ratio i4 at 4th gear is calculated as follows: Sun gear S1 rotation speed: ωS1=(1 / J12)ωi= (ωi / -1.000) =-1.000ωi Sun gear S2 rotation speed: ωS2=0 Carrier C0 rotation speed: ωC={α / (α+β)}·(1 / J12)ωi =(0.267 / 0.600)·(1 / -1.000)ωi=-0.445ωi Output speed (= ring gear R0 speed): ωo=ωR={α(1+β) / (α+β)}·(1 / J12)ωi =(0.356 / 0.600)·(1 / -1.000)ωi=-0.593ωi (See Figure 4(b)) 4th gear ratio: i4=(ωi / ωo)= (ωi / -0.593ωi) ≒-1.69

[0053] <5th gear> As shown in FIG. 7B, in 5th gear (number 5 in a circle), brakes B11 and B21 are engaged. The rotation from the input coupling mechanism IN enters the sun gear S11 of the double pinion planetary gear PG11, and because the ring gear R11 is stopped, the carrier C11 rotates. The rotation of the carrier C11 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. The reduction ratio J11 of the double pinion planetary gear PG11 is -1.500. The rotation from the input coupling mechanism IN also enters the sun gear S21 of the double pinion planetary gear PG21, and because the ring gear R21 is stopped, the carrier C21 rotates. The rotation of the carrier C21 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. The reduction ratio J21 of the double pinion planetary gear PG21 is -1.000. In other words, the rotation is input to the sun gears S1 and S2 of the Ravigneaux planetary gear PG0, forming a two-system input. The rotation of the sun gears S1 and S2 is transmitted to the ring gear R0 via the short pinion and long pinion, and the rotation of the ring gear R0 is output to the output section G0.

[0054] The gear ratio i5 at 5th gear is calculated as follows: Sun gear S1 rotation speed: ωS1=(1 / J11)ωi= (ωi / -1.500) =-0.667ωi Sun gear S2 rotation speed: ωS2=(1 / J21)ωi= (ωi / -1.000) =-1.000ωi Rotational speed of carrier C0: ωC={(α / J11)+(β / J21)} / (α+β)·ωi ={(0.267 / -1.500)+(0.333 / -1.000)} / (0.600)·ωi =(-0.178-0.333) / 0.600·ωi = -(0.665 / 0.600)·ωi =-1.108ωi Output speed (= ring gear R0 speed): ωo=ωR=[{α(1+β) / J11+β(1-α) / J21} / (α+β)]·ωi =[{(0.356 / -1.500)+(0.244 / -1.000)} / 0.600]·ωi = (-0.237-0.244) / 0.600·ωi =(-0.481 / 0.600)·ωi=-0.802ωi (see Figure 4(c)) 5th gear ratio: i5=(ωi / ωo)= (ωi / -0.802ωi) ≒-1.25

[0055] <6th gear> As shown in FIG. 7B, in 6th gear (circled number 6), brakes B12 and B21 are engaged. The rotation from the input coupling mechanism IN enters the sun gear S12 of the double pinion planetary gear PG12, and because the ring gear R12 is stopped, the carrier C12 rotates. The rotation of the carrier C12 is input to the sun gear S1 of the Ravigneaux planetary gear PG0 via the carrier C11. The reduction ratio J12 of the double pinion planetary gear PG12 is -1.000. The rotation from the input coupling mechanism IN also enters the sun gear S21 of the double pinion planetary gear PG21, and because the ring gear R21 is stopped, the carrier C21 rotates. The rotation of the carrier C21 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. The reduction ratio J21 of the double pinion planetary gear PG21 is -1.000. In other words, the rotation is input to the sun gears S1 and S2 of the Ravigneaux planetary gear PG0, forming a two-system input. The rotation of the sun gears S1 and S2 is transmitted to the ring gear R0 via the short pinion and long pinion, and the rotation of the ring gear R0 is output to the output section G0.

[0056] The gear ratio for 6th gear i6 is calculated as follows: Sun gear S1 rotation speed: ωS1=(1 / J12)ωi= (ωi / -1.000) =-1.000ωi Sun gear S2 rotation speed: ωS2=(1 / J21)ωi= (ωi / -1.000) =-1.000ωi Rotational speed of carrier C0: ωC=[{(α / J12)+(β / J21)} / (α+β)]·ωi =[{(0.267 / -1.000)+(0.333 / -1.000)} / 0.600]·ωi ={(-0.267-0.333) / 0.600}·ωi =(-0.600 / 0.600)·ωi =-1.000ωi Output speed (= ring gear R0 speed): ωo=[{α(1+β) / J12+β(1-α) / J21} / (α+β)}]·ωi =[{(0.356 / -1.000)+(0.244 / -1.000)} / 0.600]·ωi ={(-0.356-0.244) / 0.600}·ωi =(-0.600 / 0.600)·ωi=-1.000ωi (see Figure 4(c)) 6th gear ratio: i6=(ωi / ωo)= (ωi / -1.000ωi) =-1.00

[0057] In this embodiment, the planetary gears PG11, PG12, PG21, and PG0 are operated using only brakes, without using clutches. This allows for a simpler structure, smaller size, lighter weight, and lower costs, and improves the freedom of layout of the entire AT. In addition, control is simple because at each gear position, each brake is either in a braking state or a non-braking state according to the brake application table.

[0058] In this embodiment, the structure is simplified by braking the ring gear in each of the double pinion planetary gears B11, B12, B21. Since the ring gear is located on the outside close to the case, a configuration in which the ring gear is braked is advantageous in terms of layout.

[0059] In this embodiment, two shift speeds are input to the sun gear S1 of the Ravigneaux planetary gear PG0 (two-system input). That is, the carrier C11 of the double pinion planetary gear PG11 and the carrier C12 of the double pinion planetary gear PG12 are connected to the sun gear S1 (input member) of the Ravigneaux planetary gear PG0. Then, the input from the double pinion planetary gear PG11 and the input from the double pinion planetary gear PG12 are switched so that either one is input to the sun gear S1 by the operation of the brakes B11 and B12. That is, since a plurality of shift speeds are input to one input member of the compound planetary gear, the gear ratio can be increased with a simple structure.

[0060] <Second embodiment> A speed change mechanism FA0603 for a stepped automatic transmission for an automobile according to a second embodiment of the present invention will be described with reference to Figs. 8A and 8B. Fig. 8A is a skeleton diagram of a speed change mechanism FA0603 for a front-wheel drive 6-speed forward / 3-speed reverse automatic transmission. To avoid complexity, the rotating body is illustrated on one side (upper half) with the rotation axis X as the center.

[0061] The speed change mechanism FA0603 includes seven double pinion planetary gears PGR1, PGR2, PG13, PG46, PGA, PGB, and PGC. Each of the double pinion planetary gears PGR1, PGR2, PG13, PG46, PGA, PGB, and PGC has a known configuration. As shown in Fig. 8A, a main gear train including the double pinion planetary gears PGR1, PGR2, PG13, PG46, PGA, PGB, and PGC is disposed on one shaft.

[0062] Double pinion planetary gear PGR1 includes a sun gear SR1, a ring gear RR1, and a carrier CR1. Double pinion planetary gear PGR2 includes a sun gear SR2, a ring gear RR2, and a carrier CR2. Double pinion planetary gear PG13 includes a sun gear S13, a ring gear R13, and a carrier C13. Double pinion planetary gear PG46 includes a sun gear S46, a ring gear R46, and a carrier C46. Double pinion planetary gear PGA includes a sun gear SA, a ring gear RA, and a carrier CA. Double pinion planetary gear PGB includes a sun gear SB, a ring gear RB, and a carrier CB. Double pinion planetary gear PGC includes a sun gear SC, a ring gear RC, and a carrier CC.

[0063] The tooth ratio of the double pinion planetary gear PGR1 is λR1 = 21 / 90 = 0.233. The tooth ratio of the double pinion planetary gear PGR2 is λR2 = 45 / 90 = 0.500. The tooth ratio of the double pinion planetary gear PG13 is λ13 = 21 / 90 = 0.233. The tooth ratio of the double pinion planetary gear PG46 is λ46 = 45 / 90 = 0.500. The tooth ratio of the double pinion planetary gear PGA is λA = 36 / 90 = 0.400. The tooth ratio of the double pinion planetary gear PGB is λB = 45 / 90 = 0.500. The tooth ratio of the double pinion planetary gear PGC is λC = 36 / 90 = 0.400.

[0064] The transmission mechanism FA0603 further includes brakes BR1, BR2, B13, B46, BA, BB, and BC. Each of these brakes is a wet multi-plate type in which rotating discs (the side to be braked) and fixed discs are packed alternately, and the fixed discs are fitted into the case K. The brake BR1 stops the rotation of the ring gear RR1 of the double pinion planetary gear PGR1. The brake BR2 stops the rotation of the ring gear RR2 of the double pinion planetary gear PGR2. The brake B13 stops the rotation of the ring gear R13 of the double pinion planetary gear PG13. The brake B46 stops the rotation of the ring gear R46 of the double pinion planetary gear PG46. The brake BA stops the rotation of the ring gear RA of the double pinion planetary gear PGA. The brake BB stops the rotation of the ring gear RB of the double pinion planetary gear PGB. Brake BC stops the rotation of ring gear RC of double pinion planetary gear PGC.

[0065] The input coupling mechanism IN is the same as that described above. The input rotation from the input coupling mechanism IN is input to the sun gear SA of the double pinion planetary gear PGA, the sun gear SB of the double pinion planetary gear PGB, and the carrier CC of the double pinion planetary gear PGC.

[0066] The speed change mechanism FA0603 is connected to an output section G10. The output section G10 includes a speed change mechanism output gear G11 and an output gear set G12. The output gear set G12 drives a differential device (not shown). The differential device (not shown) includes a final reduction gear.

[0067] The rotation of the carrier C46 of PG46 is transmitted to the speed change mechanism output gear G11, and this rotation is in turn transmitted to the output gear set G12 and the differential device (not shown).

[0068] <At reverse 1st gear> As shown in FIG. 8B, in first reverse gear (R1 (number 1 in a circle)), brakes BR1, BR2, and BA are engaged. Rotation from the input coupling mechanism IN enters the sun gear SA of the double pinion planetary gear PGA, and because the ring gear RA is stopped, the carrier CA rotates. The rotation of the carrier CA is input to the sun gear SR1 of the double pinion planetary gear PGR1. Because the ring gear RR1 is stopped, the rotation of the sun gear SR1 is transmitted to the carrier CR1, and further to the sun gear SR2 of the double pinion planetary gear PGR2. Because the ring gear RR2 is stopped, the rotation of the sun gear SR2 is transmitted to the carrier CR2. Because the brakes B13 and B46 are not engaged, the carriers C13 and C46 of the double pinion planetary gears PG13 and PG46 rotate in the same manner as the carrier CR2 of the double pinion planetary gear PGR2. The rotation of the carrier C46 is output to the output section G10.

[0069] The gear ratio iR1 in first reverse gear is calculated as follows: Here, the driving force is transmitted in the order of double pinion planetary gears PGA, PGR1, and PGR2. For the double pinion planetary gear PGA, the ring gear rotation speed ωR = 0, so from the above equation (2), the carrier rotation speed ωC = (-λ) / (1-λ) ωS, and the reduction ratio JA = (1-λ) / (-λ) = (1-36 / 90) / (-36 / 90) = -1.500. Calculating in the same way, the reduction ratio JR1 of the double pinion planetary gear PGR1 is -3.292, and the reduction ratio JR2 of the double pinion planetary gear PGR2 is -1.000. Therefore, the gear ratio at 1st reverse gear is: iR1 = JA × JR1 × JR2 = (-1.500) × (-3.292) × (-1.000) ≒ -4.94.

[0070] <2nd reverse gear> As shown in Figure 8B, in second reverse gear (R2), brakes BR1, BR2, and BB are engaged. Rotation from the input coupling mechanism IN enters sun gear SB of double pinion planetary gear PGB, and because ring gear RB is stopped, carrier CB rotates. Carrier CA also rotates together with carrier CB. The rotation of carrier CA is input to sun gear SR1 of double pinion planetary gear PGR1. The following explanation is the same as for first reverse gear.

[0071] The gear ratio iR2 in second reverse gear is calculated as follows: Here, the driving force is transmitted in the order of double pinion planetary gears PGB, PGR1, and PGR2. For the double pinion planetary gear PGB, the ring gear rotation speed ωR = 0, so from the above equation (2), the carrier rotation speed ωC = (-λ) / (1-λ) ωS, and the reduction ratio JB = (1-λ) / (-λ) = (1-45 / 90) / (-45 / 90) = -1.000. Therefore, the gear ratio for reverse 2nd gear is: iR2=JB×JR1×JR2=(-1.000)×(-3.292)×(-1.000)≒-3.29.

[0072] <3rd reverse gear> As shown in Figure 8B, in third reverse gear (R3), brakes BR1, BR2, and BC are engaged. Rotation from the input coupling mechanism IN enters carrier CC of double pinion planetary gear PGC, and because ring gear RC is stopped, sun gear SC rotates. Carriers CB and CA of double pinion planetary gears PGB and PGA rotate together with the sun gear SC, and the rotation of carrier CA is input to sun gear SR1 of double pinion planetary gear PGR1. The following explanation is the same as for first reverse gear.

[0073] The gear ratio iR3 in third reverse gear is calculated as follows: Here, the driving force is transmitted in the order of double pinion planetary gears PGC, PGR1, and PGR2. For the double pinion planetary gear PGC, the ring gear rotation speed ωR = 0, so from the above equation (2), the sun gear rotation speed ωS = (1-λ) / (-λ) ωC, and the reduction ratio JC = (-λ) / (1-λ) = (-36 / 90) / (1-36 / 90) = -0.667. Therefore, the gear ratio for reverse 3 is: iR3 = JC × JR1 × JR2 = (-0.667) × (-3.292) × (-1.000) ≒ -2.20.

[0074] <1st gear> As shown in FIG. 8B, in first gear, brakes B13 and BA are engaged. Rotation from the input coupling mechanism IN enters the sun gear SA of the double pinion planetary gear PGA, and because the ring gear RA is stopped, the carrier CA rotates. The rotation of the carrier CA is input to the sun gear S13 of the double pinion planetary gear PG13. Because the ring gear R13 is stopped, the rotation of the sun gear S13 is transmitted to the carrier C13. The carrier C46 of the double pinion planetary gear PG46 rotates together with the carrier C13 of the double pinion planetary gear PG13. The rotation of the carrier C46 is output to the output section G10. Thus, during forward travel, the brakes BR1 and BR2 are not engaged for the reverse double pinion planetary gears PGR1 and PGR2, and are in a completely free state.

[0075] The gear ratio i1 in 1st gear is calculated as follows: Here, the driving force is transmitted to the double pinion planetary gears PGA and PG13 in that order. For double pinion planetary gear PGA, the reduction ratio JA = (1-λ) / (-λ) = (1-36 / 90) / (-36 / 90) = -1.500 as described above. For double pinion planetary gear PG13, since the ring gear rotation speed ωR = 0, the carrier rotation speed ωC = (-λ) / (1-λ) ωS from the above formula (2) becomes, and the reduction ratio J13 = (1-λ) / (-λ) = (1-21 / 90) / (-21 / 90) = -3.286. Therefore, the gear ratio at 1st gear is i1 = JA × J13 = (-1.500) × (-3.286) ≒ 4.94.

[0076] <2nd gear> As shown in Figure 8B, in second gear (number 2 in a circle), brakes B13 and BB are engaged. Rotation from the input coupling mechanism IN enters the sun gear SB of the double pinion planetary gear PGB, and because the ring gear RB is stopped, the carrier CB rotates. Carrier CA rotates together with carrier CB, and the rotation of carrier CA is input to the sun gear S13 of the double pinion planetary gear PG13. The rest of the process is the same as in first gear.

[0077] The gear ratio i2 in second gear is calculated as follows. Here, the driving force is transmitted to the double pinion planetary gears PGB and PG13 in that order. For the double pinion planetary gear PGB, the reduction ratio JB is -1.000 as described above. For the double pinion planetary gear PG13, the reduction ratio J13 is -3.286 as described above. Therefore, the gear ratio at 2nd gear is i2 = JB × J13 = (-1.000) × (-3.286) ≒ 3.29.

[0078] <3rd gear> As shown in Figure 8B, in third gear (circled number 3), brakes B13 and BC are engaged. Rotation from the input coupling mechanism IN enters carrier CC of double pinion planetary gear PGC, and because ring gear RC is stopped, sun gear SC rotates. Sun gear SC rotates together with carriers CB and CA of double pinion planetary gears PGB and PGA, and the rotation of carrier CA is input to sun gear S13 of double pinion planetary gear PG13. The rest of the process is the same as in first gear.

[0079] The gear ratio i3 in third gear is calculated as follows: Here, the driving force is transmitted to the double pinion planetary gears PGC and PG13 in that order. For the double pinion planetary gear PGC, the reduction ratio JC is -0.667 as described above. For the double pinion planetary gear PG13, the reduction ratio J13 is -3.286 as described above. Therefore, the gear ratio at 3rd gear is i2 = JC × J13 = (-0.667) × (-3.286) ≒ 2.20.

[0080] <4th gear> As shown in FIG. 8B, in fourth gear (number 4 in a circle), brakes B46 and BA are engaged. Rotation from the input coupling mechanism IN enters the sun gear SA of the double pinion planetary gear PGA, and because the ring gear RA is stopped, the carrier CA rotates. The rotation of carrier CA is input to the sun gear S46 of the double pinion planetary gear PG46. Because the ring gear R46 is stopped, the rotation of the sun gear S46 is transmitted to the carrier C46. The rotation of carrier C46 is output to the output section G10.

[0081] The gear ratio i4 in 4th gear is calculated as follows. Here, the driving force is transmitted to the double pinion planetary gear PGA, then to PG46. For the double pinion planetary gear PGA, the reduction ratio JA = -1.500 as described above. For the double pinion planetary gear PG46, the ring gear rotation speed ωR = 0, so from the above formula (2), the carrier rotation speed ωC = (-λ) / (1-λ) ωS, and the reduction ratio J46 = (1-λ) / (-λ) = (1-45 / 90) / (-45 / 90) = -1.000. Therefore, the gear ratio at 4th gear is i4=JA×J46 = (-1.500)×(-1.000)≒1.50.

[0082] <5th gear> As shown in FIG. 8B, in 5th gear (number 5 in a circle), brakes B46 and BB are engaged. Rotation from the input coupling mechanism IN enters sun gear SB of double pinion planetary gear PGB, and because ring gear RB is stopped, carrier CB rotates. Carrier CA of double pinion planetary gear PGA rotates together with carrier CB, and the rotation of carrier CA is input to sun gear S46 of double pinion planetary gear PG46. The rest of the process is the same as in 4th gear.

[0083] The gear ratio i5 in 5th gear is calculated as follows. Here, the driving force is transmitted to the double pinion planetary gear PGB, then to PG46. For the double pinion planetary gear PGB, the reduction ratio JB = -1.000 as described above.For the double pinion planetary gear PG46, the reduction ratio J46 = -1.000 as described above. Therefore, the gear ratio at 5th gear is i5=JB×J46 = (-1.000)×(-1.000)≒1.00.

[0084] <6th gear> As shown in Figure 8B, in sixth gear (number 6 in a circle), brakes B46 and BC are engaged. Rotation from the input coupling mechanism IN enters carrier CC of double pinion planetary gear PGC, and because ring gear RC is stopped, sun gear SC rotates. Sun gear SC rotates together with carriers CB and CA of double pinion planetary gears PGB and PGA, and the rotation of carrier CA is input to sun gear S46 of double pinion planetary gear PG46. The rest of the process is the same as in fourth gear.

[0085] The gear ratio i6 in 6th gear is calculated as follows. Here, the driving force is transmitted in the order of double pinion planetary gear PGC and PG46. For the double pinion planetary gear PGC, the reduction ratio JC is -0.667 as described above. For the double pinion planetary gear PG46, the reduction ratio J46 is -1.000 as described above. Therefore, the gear ratio at 6th gear is i6=JC×J46 = (-0.667)×(-1.000)≒0.67.

[0086] <Functions of each planetary gear group> In the transmission mechanism FA0603, the double pinion planetary gears PGA, PGB, PGC form a first planetary gear group G1, the double pinion planetary gears PGR1, PGR2 form a second planetary gear group G2, and the double pinion planetary gears PG13, PG46 form a third planetary gear group G3.

[0087] For the first planetary gear group G1, as shown in FIG. 8B, at each gear position, one of the double pinion planetary gears PGA, PGB, PGC is operated by engaging one of the brakes BA, BB, BC. The double pinion planetary gears PGA and PGB are both sun gear input / carrier output, and the double pinion planetary gear PGC is carrier input / sun gear output. Therefore, regardless of which of the double pinion planetary gears PGA, PGB, PGC is operating, the output rotation is reversed relative to the input. Therefore, the first planetary gear group G1 as a group has a reverse output function.

[0088] As shown in FIG. 8B, in the second planetary gear group G2, the brakes BR1 and BR2 are both engaged in reverse 1 to 3 gears, so that the double pinion planetary gears PGR1 and PGR2 are both operated. The double pinion planetary gears PGR1 and PGR2 both have a sun gear input and a carrier output, and the output rotation is reversed relative to the input. Since the output of the double pinion planetary gear PGR1 is the input of the double pinion planetary gear PGR2, it can be said that the double pinion planetary gears PGR1 and PGR2 are connected in series. Therefore, when the double pinion planetary gears PGR1 and PGR2 are viewed as a group, the input rotation direction and the output rotation direction are the same (reverse + reverse = forward). In other words, the second planetary gear group G2 has a forward rotation (same direction rotation) output function as a group. That is, when viewed as the second planetary gear group G2, an output rotation in the same direction as the input rotation direction is obtained.

[0089] As shown in FIG. 8B, for the third planetary gear group G3, in the first to third forward gears, the brake B13 is engaged to operate the double pinion planetary gear PG13, and in the fourth to sixth forward gears, the brake B46 is engaged to operate the double pinion planetary gear PG46. Both the double pinion planetary gears PG13 and PG46 are sun gear input / carrier output. Therefore, whichever of the double pinion planetary gears PG13 and PG46 is in operation, the output rotation is reversed relative to the input. Therefore, the third planetary gear group G3 as a group has a reverse output function.

[0090] In the above configuration, the first planetary gear group G1 and the third planetary gear group G3 connected in series are used when traveling forward. As described above, the first planetary gear group G1 and the third planetary gear group G3 both have a reverse rotation output function, so the transmission outputs forward rotation (reverse rotation + reverse rotation = forward rotation). When traveling backward, the first planetary gear group G1 and the second planetary gear group G2 connected in series are used. As described above, the first planetary gear group G1 has a reverse rotation output function, and the second planetary gear group G2 has a forward rotation output function, so the transmission outputs reverse rotation (reverse rotation + forward rotation = reverse rotation).

[0091] <Effects> In this embodiment, the planetary gears PGA, PGB, PGC, PG13, PG46, PGR1, and PGR2 are operated using only brakes, without using clutches, which makes it possible to simplify the structure, reduce the size and weight, and reduce costs, thereby improving the freedom of layout of the entire AT.

[0092] In this embodiment, the structure is simplified by braking the ring gear in each of the double pinion planetary gears PGA, PGB, PGC, PG13, PG46, PGR1, and PGR2. Since the ring gear is located on the outside close to the case, a configuration in which the ring gear is braked is advantageous in terms of layout.

[0093] In addition, since this embodiment has multiple reverse gears, it is possible to use a gear with an appropriate driving force even when reversing. Furthermore, since it uses only brakes and no clutches, the overall structure of the AT is compact and there is a large degree of freedom in layout, so it is easy to realize multiple reverse gears in terms of layout.

[0094] In this embodiment, two planetary gear groups (G1, G3) each having a reverse output function are connected in series to provide a forward rotation output as a transmission. This allows output rotation in the same direction as the input rotation direction to be obtained when moving forward. In addition, a first planetary gear group G1 having a reverse output function and a second planetary gear group G2 having a forward output function are connected in series to provide a reverse rotation output as a transmission. This allows output rotation in the opposite direction to the input rotation direction to be obtained when moving backward.

[0095] <Third embodiment> A shift mechanism RB0803 for a stepped automatic transmission for an automobile according to a third embodiment of the present invention will be described with reference to Figs. 9A and 9B. Fig. 9A is a skeleton diagram of a shift mechanism for a rear-wheel drive 8-speed forward / 2-speed reverse automatic transmission. To avoid complication, the rotating body is illustrated on one side (upper half) with the rotation axis X as the center.

[0096] The speed change mechanism RB0803 includes five sets of double pinion planetary gears PGZ, PG11, PG12, PG21, PG22 and a Ravigneaux planetary gear PG0.

[0097] The double pinion planetary gears PGZ, PG11, PG12, PG21, PG22 and the Ravigneaux planetary gear PG0 each have a known configuration. The double pinion planetary gear PGZ includes a sun gear SZ, a ring gear RZ, and a carrier CZ. The double pinion planetary gear PG11 includes a sun gear S11, a ring gear R11, and a carrier C11. The double pinion planetary gear PG12 includes a sun gear S12, a ring gear R12, and a carrier C12. The double pinion planetary gear PG21 includes a sun gear S21, a ring gear R21, and a carrier C21. The double pinion planetary gear PG22 includes a sun gear S22, a ring gear R22, and a carrier C22. The Ravigneaux planetary gear PG0 includes two sun gears S1 and S2, a ring gear R0, and a carrier C0.

[0098] The tooth ratio of the double pinion planetary gear PGZ is λZ = 45 / 90 = 0.500. The tooth ratio of the double pinion planetary gear PG11 is λ11 = 60 / 105 = 0.571. The tooth ratio of the double pinion planetary gear PG12 is λ12 = 42 / 90 = 0.467. The tooth ratio of the double pinion planetary gear PG21 is λ21 = 45 / 90 = 0.500. The tooth ratio of the double pinion planetary gear PG22 is λ22 = 39 / 90 = 0.433. The tooth ratio of the sun gear S1 of the Ravigneaux planetary gear PG0 is α = 24 / 90 = 0.267, and the tooth ratio of the sun gear S2 of the Ravigneaux planetary gear PG0 is β = 30 / 90 = 0.333.

[0099] The transmission mechanism RB0803 further includes a ring gear fixing member Z and brakes B11, B12, B21, B22, B02, and B01. The ring gear fixing member Z always stops the ring gear RZ of the double pinion planetary gear PGZ, and can be considered a brake in a broad sense. Each of the brakes B11, B12, B21, B22, B02, and B01 is a wet multi-plate type in which the rotating side disc (the side to be braked) and the fixed side disc are alternately packed, and the fixed side disc is fitted into the case K. The brake B11 stops the rotation of the ring gear R11 of the double pinion planetary gear PG11. The brake B12 stops the rotation of the ring gear R12 of the double pinion planetary gear PG12. The brake B21 stops the rotation of the ring gear R21 of the double pinion planetary gear PG21. The brake B22 stops rotation of the ring gear R22 of the double pinion planetary gear PG22. The brake B01 stops rotation of the carrier C0 of the Ravigneaux planetary gear PG0. The brake B02 stops rotation of the sun gear S21 of the double pinion planetary gear PG21, the sun gear S22 of the double pinion planetary gear PG22, and the sun gear S2 of the Ravigneaux planetary gear PG0.

[0100] The input coupling mechanism IN is a fluid coupling, a torque converter, a wet clutch, or the like. The input rotation from the input coupling mechanism IN is input to the sun gear SZ of the double pinion planetary gear PGZ. The speed change mechanism RB0803 also has an output section (OUT). The rotation of the ring gear R0 of the Ravigneaux planetary gear PG0 is transmitted to an output gear set (not shown).

[0101] <At reverse 1st gear> As shown in FIG. 9B, in first reverse gear (R1), the brakes B21 and B01 are engaged. The rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and because the ring gear RZ is stopped, the carrier CZ rotates. The carrier C21 of the double pinion planetary gear PG21 rotates together with the carrier CZ, carrier C11, and carrier C12, and because the brake B21 is engaged, the rotation is transmitted to the sun gear S21. The rotation of the sun gear S21 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. Because the carrier C0 is stopped, the rotation of the sun gear S2 is transmitted to the ring gear R0. The rotation of the ring gear R0 is output to the output unit OUT.

[0102] For the double pinion planetary gear PGZ, the ring gear rotation speed ωR = 0, so from the above (Equation 2), the carrier rotation speed ωC = -1.000ωS, and the reduction ratio JZ = -1.000. In other words, the double pinion planetary gear PGZ at the most upstream inputs a constant reverse speed output to the downstream speed change function unit. Similarly, for the double pinion planetary gear PG21, the carrier rotation speed ωC = -1.000ωS, and the reduction ratio J21 = -1.000. For the Ravigneaux planetary gear PG0, the carrier is fixed (ωC = 0) in the above equation (3B), so ωR = -β·ωS2 = (-30 / 90)·ωS2 = (-0.333)·ωS2 So the reduction ratio is J0 = -3.000. Therefore, the gear ratio for reverse 1st gear is iR1 = JZ × J21 × J0 = (-1.000) × (-1.000) × (-3.000) ≒ -3.00.

[0103] <2nd reverse gear> As shown in FIG. 9B, in second reverse gear (R2), the brakes B22 and B01 are engaged. The rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and because the ring gear RZ is stopped, the carrier CZ rotates. The carrier C22 of the double pinion planetary gear PG22 rotates together with the carrier CZ, carrier C11, carrier C12, and carrier C21, and because the brake B22 is engaged, the rotation is transmitted to the sun gear S22. The rotation of the sun gear S22 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. Because the carrier C0 is stopped, the rotation of the sun gear S2 is transmitted to the ring gear R0. The rotation of the ring gear R0 is output to the output unit OUT.

[0104] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as above. For the double pinion planetary gear PG22, from the above (Equation 2), (1-39 / 90) · ωC = (-39 / 90) · ωS, so the reduction ratio J22 = -0.765. For the Ravigneaux planetary gear PG0, the reduction ratio J0 = -3.000, as above. Therefore, the gear ratio for second reverse gear is, iR2 = JZ × J22 × J0 = (-1.000) × (-0.765) × (-3.000) ≒ -2.29.

[0105] <First forward gear> As shown in FIG. 9B, in first forward gear (number 1 in a circle), brakes B11 and B01 are engaged. Rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and because the ring gear RZ is stopped, the carrier CZ rotates. The carrier C11 of the double pinion planetary gear PG11 rotates together with the carrier CZ, and because the brake B11 is engaged, the rotation is transmitted to the sun gear S11. The rotation of the sun gear S11 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Because the carrier C0 is stopped, the rotation of the sun gear S1 is transmitted to the ring gear R0. The rotation of the ring gear R0 is output to the output section OUT.

[0106] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as above. For the double pinion planetary gear PG11, from the above (Equation 2), (1-60 / 105) · ωC = (-60 / 105) · ωS, so the reduction ratio J11 = -1.333. For the Ravigneaux planetary gear PG0, since the carrier is fixed (ωC = 0) in the above equation (3A), ωR = α · ωS1 = (24 / 90) · ωS1 = (0.267) · ωS1, and the reduction ratio J0 = 3.750. Therefore, the gear ratio at 1st forward speed is, i1 = JZ × J11 × J0 = (-1.000) × (-1.333) × (3.750) ≒ 5.00.

[0107] <2nd forward gear> As shown in FIG. 9B, in second forward gear (number 2 in a circle), brakes B12 and B01 are engaged. Rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and because the ring gear RZ is stopped, the carrier CZ rotates. The carrier C12 of the double pinion planetary gear PG12 rotates together with the carrier CZ and carrier C11, and because the brake B12 is engaged, the rotation is transmitted to the sun gear S12. The rotation of the sun gear S12 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Because the carrier C0 is stopped, the rotation of the sun gear S1 is transmitted to the ring gear R0. The rotation of the ring gear R0 is output to the output section OUT.

[0108] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as above. For the double pinion planetary gear PG12, from the above (Equation 2), (1-42 / 90) · ωC = (-42 / 90) · ωS, so the reduction ratio J12 = -0.875. For the Ravigneaux planetary gear PG0, the reduction ratio J0 = 3.750, as above. Therefore, the gear ratio at second forward speed is, i2 = JZ × J12 × J0 = (-1.000) × (-0.875) × (3.750) ≒ 3.28.

[0109] <Forward 3rd gear> As shown in FIG. 9B, in third forward gear (number 3 in a circle), brakes B11 and B02 are engaged. Rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and because the ring gear RZ is stopped, the carrier CZ rotates. The carrier C11 of the double pinion planetary gear PG11 rotates together with the carrier CZ, and because the brake B11 is engaged, the rotation is transmitted to the sun gear S11. The rotation of the sun gear S11 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Because the sun gear S2 is stopped, the rotation of the sun gear S1 is transmitted to the ring gear R0 via the carrier C0. The rotation of the ring gear R0 is output to the output section OUT.

[0110] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as above. For the double pinion planetary gear PG11, the reduction ratio J11 = -1.333, as above. For the Ravigneaux planetary gear PG0, the input is the sun gear S1, the sun gear S2 is fixed, and the output is the ring gear R0. Therefore, if ωS2 = 0 in the above equations (3A) and (3B), then ωR = α(1 + β) / (α + β) · ωS1, and the reduction ratio J0 = (α + β) / {α(1 + β)} = 1.688. Note that (b) in Figure 4 may also be used directly. Therefore, the gear ratio at third forward speed is, i3 = JZ × J11 × J0 = (-1.000) × (-1.333) × (1.688) ≒ 2.25.

[0111] <Forward 4th gear> As shown in FIG. 9B, in fourth forward gear (number 4 in a circle), brakes B12 and B02 are engaged. Rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and because the ring gear RZ is stopped, the carrier CZ rotates. The carrier C12 of the double pinion planetary gear PG12 rotates together with the carriers CZ and C11, and because the brake B12 is engaged, the rotation is transmitted to the sun gear S12. The rotation of the sun gear S12 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Because the sun gear S2 is stopped, the rotation of the sun gear S1 is transmitted to the ring gear R0 via the carrier C0. The rotation of the ring gear R0 is output to the output section OUT.

[0112] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as above. For the double pinion planetary gear PG12, the reduction ratio J12 = -0.875, as above. For the Ravigneaux planetary gear PG0, the reduction ratio J0 = 1.688, as above. Therefore, the gear ratio at 4th forward speed is: i4=JZ×J12×J0=(-1.000)×(-0.875)×(1.688)≒1.48.

[0113] <Forward 5th gear> As shown in FIG. 9B, in the fifth forward gear (number 5 in a circle), the brakes B11 and B21 are engaged. The rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and since the ring gear RZ is stopped, the carrier CZ rotates. The carrier C11 of the double pinion planetary gear PG11 rotates together with the carrier CZ, and since the brake B11 is engaged, the rotation is transmitted to the sun gear S11. The rotation of the sun gear S11 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Furthermore, the carrier C21 of the double pinion planetary gear PG21 rotates together with the carrier CZ, carrier C11, and carrier C12, and since the brake B21 is engaged, the rotation is transmitted to the sun gear S21. The rotation of the sun gear S21 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. In other words, two systems (sun gears S1, S2) are input to the Ravigneaux planetary gear PG0. The rotation of the sun gears S1 and S2 is transmitted to the ring gear R0 via the carrier C0. The rotation of the ring gear R0 is output to the output section OUT.

[0114] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as described above. For the double pinion planetary gear PG11, the reduction ratio J11 = -1.333, as described above. For the double pinion planetary gear PG21, the reduction ratio J21 = -1.000, as described above. For the Ravigneaux planetary gear PG0, by eliminating ωC from the above equations (3A) and (3B) and solving for ωR, ωR = {α(1+β)ωS1 + β(1-α) · ωS2} / (α+β), and the reduction ratio J0 = (α+β) / [{α(1+β) / J1} + {β(1-α) / J2}] is obtained. Here, J1 is the reduction ratio of the sun gear S1 input, and J2 is the reduction ratio of the sun gear S2 input. Note that (c) in FIG. 4 may also be used. Therefore, the gear ratio for 5th forward gear is i5= JZ×J0 = (-1.000)×(α+β) / [{α(1+β) / (-1.333)}+{β(1-α) / (-1.000)}]≒1.17.

[0115] <Forward 6th gear> As shown in FIG. 9B, in the sixth forward speed (circled number 6), the brakes B11 and B22 are engaged. The rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and since the ring gear RZ is stopped, the carrier CZ rotates. The carrier C11 of the double pinion planetary gear PG11 rotates together with the carrier CZ, and since the brake B11 is engaged, the rotation is transmitted to the sun gear S11. The rotation of the sun gear S11 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Furthermore, the carrier C22 of the double pinion planetary gear PG22 rotates together with the carrier CZ, carrier C11, carrier C12, and carrier C21, and since the brake B22 is engaged, the rotation is transmitted to the sun gear S22. The rotation of the sun gear S22 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. The rotation of the sun gears S1 and S2 is transmitted to the ring gear R0 via the carrier C0. The rotation of the ring gear R0 is output to the output section OUT.

[0116] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as above. For the double pinion planetary gear PG11, the reduction ratio J11 = -1.333, as above. For the double pinion planetary gear PG22, the reduction ratio J22 = -0.765, as above. For the Ravigneaux planetary gear PG0, the reduction ratio J0 = (α + β) / [{α(1 + β) / J1} + {β(1 - α) / J2}], as above. Therefore, the gear ratio at 6th forward speed is: i6= JZ×J0 = (-1.000)×(α+β) / [{α(1+β) / (-1.333)}+{β(1-α) / (-0.765)}]≒1.02.

[0117] <Forward 7th gear> As shown in FIG. 9B, in the seventh forward speed (number 7 in a circle), the brakes B12 and B21 are engaged. The rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and since the ring gear RZ is stopped, the carrier CZ rotates. The carrier C12 of the double pinion planetary gear PG12 rotates together with the carrier CZ and carrier C11, and since the brake B12 is engaged, the rotation is transmitted to the sun gear S12. The rotation of the sun gear S12 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Furthermore, the carrier C21 of the double pinion planetary gear PG21 rotates together with the carrier CZ, carrier C11, and carrier C12, and since the brake B21 is engaged, the rotation is transmitted to the sun gear S21. The rotation of the sun gear S21 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. The rotation of the sun gears S1 and S2 is transmitted to the ring gear R0 via the carrier C0. The rotation of the ring gear R0 is output to the output section OUT.

[0118] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as above. For the double pinion planetary gear PG12, the reduction ratio J12 = -0.875, as above. For the double pinion planetary gear PG21, the reduction ratio J21 = -1.000, as above. For the Ravigneaux planetary gear PG0, the reduction ratio J0 = (α + β) / [{α(1 + β) / J1} + {β(1 - α) / J2}], as above. Therefore, the gear ratio at 7th forward speed is: i7= JZ×J0 = (-1.000)×(α+β) / [{α(1+β) / (-0.875)}+{β(1-α) / (-1.000)}]≒0.92.

[0119] <Forward 8th gear> As shown in FIG. 9B, in the eighth forward speed (number 8 in a circle), the brakes B12 and B22 are engaged. The rotation from the input coupling mechanism IN enters the sun gear SZ of the double pinion planetary gear PGZ, and since the ring gear RZ is stopped, the carrier CZ rotates. The carrier C12 of the double pinion planetary gear PG12 rotates together with the carrier CZ and carrier C11, and since the brake B12 is engaged, the rotation is transmitted to the sun gear S12. The rotation of the sun gear S12 is input to the sun gear S1 of the Ravigneaux planetary gear PG0. Furthermore, the carrier C22 of the double pinion planetary gear PG22 rotates together with the carrier CZ, carrier C11, carrier C12, and carrier C21, and since the brake B22 is engaged, the rotation is transmitted to the sun gear S22. The rotation of the sun gear S22 is input to the sun gear S2 of the Ravigneaux planetary gear PG0. The rotation of the sun gears S1 and S2 is transmitted to the ring gear R0 via the carrier C0. The rotation of the ring gear R0 is output to the output section OUT.

[0120] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000, as above. For the double pinion planetary gear PG12, the reduction ratio J12 = -0.875, as above. For the double pinion planetary gear PG22, the reduction ratio J22 = -0.765, as above. For the Ravigneaux planetary gear PG0, the reduction ratio J0 = (α + β) / [{α(1 + β) / J1} + {β(1 - α) / J2}], as above. Therefore, the gear ratio at 8th forward speed is: i8= JZ×J0 = (-1.000)×(α+β) / [{α(1+β) / (-0.875)}+{β(1-α) / (-0.765)}]≒0.83.

[0121] In this embodiment, there are two inputs to the sun gear S1 of the Ravigneaux planetary gear PG0 (S1×2). That is, the input from the double pinion planetary gear PG11 and the input from the double pinion planetary gear PG12 are switched by the operation of the brakes B11 and B12. There are also two inputs to the sun gear S2 of the Ravigneaux planetary gear PG0 (S2×2). That is, the input from the double pinion planetary gear PG21 and the input from the double pinion planetary gear PG22 are switched by the operation of the brakes B21 and B22. There is no input to the carrier C0 of the Ravigneaux planetary gear PG0.

[0122] <Effects> In this embodiment, the planetary gears PGZ, PG11, PG12, PG21, PG22, and PG0 are operated using only brakes, without using clutches. This allows for a simplified structure, smaller size, lighter weight, and reduced costs, improving the freedom of layout of the entire AT. In addition, since there are multiple reverse gears, it is possible to use a gear with an appropriate driving force even when reversing. Furthermore, since there are no clutches and only brakes are used, the structure of the entire AT is compact and there is a large degree of freedom in layout, making it easy to realize multiple reverse gears in terms of layout.

[0123] In this embodiment, two shift speeds are input to the sun gear S1 of the Ravigneaux planetary gear PG0, and two shift speeds are also input to the sun gear S2. That is, the sun gear S11 of the double pinion planetary gear PG11 and the sun gear S12 of the double pinion planetary gear PG12 are connected to the sun gear S1 (input member) of the Ravigneaux planetary gear PG0. Then, the input from the double pinion planetary gear PG11 and the input from the double pinion planetary gear PG12 are switched so that either one is input to the sun gear S1 by the operation of the brakes B11 and B12. Furthermore, the sun gear S21 of the double pinion planetary gear PG21 and the sun gear S22 of the double pinion planetary gear PG22 are connected to the sun gear S2 (input member) of the Ravigneaux planetary gear PG0. The input from double pinion planetary gear PG21 and the input from double pinion planetary gear PG22 are switched so that either one is input to sun gear S2 by operation of brakes B21, B22. That is, since a plurality of shift speeds are input to each of the two input members (sun gears S1, S2) of the compound planetary gear, the gear ratio can be increased with a simple structure.

[0124] <Fourth embodiment> A speed change mechanism FA0501 of a stepped automatic transmission for an automobile according to a fourth embodiment of the present invention will be described with reference to Figs. 10A and 10B. Fig. 10A is a skeleton diagram of a speed change mechanism of a front-wheel drive 5-speed forward / 5-speed reverse automatic transmission. To avoid complexity, the rotating body is illustrated on one side (upper half) with the rotation axis X as the center.

[0125] The speed change mechanism FA0501 includes six double pinion planetary gears PGF, PG1, PG2, PG3, PG4, and PG5, and a clutch hub sleeve CHS. The double pinion planetary gears PGF, PG1, PG2, PG3, PG4, and PG5, and the clutch hub sleeve CHS each have a known configuration. A main gear train consisting of the double pinion planetary gears PGF, PG1, PG2, PG3, PG4, and PG5, and the clutch hub sleeve CHS is disposed on one axis (rotation axis X). The double pinion planetary gear PGF includes a sun gear SF, a ring gear RF, and a carrier CF. The double pinion planetary gear PG1 includes a sun gear S1, a ring gear R1, and a carrier C1. The same is true for the double pinion planetary gears PG2, PG3, PG4, and PG5. The clutch hub sleeve CHS includes a hub H and a sleeve SL. The double pinion planetary gear PGF has a tooth ratio λF = 45 / 90 = 0.500. The double pinion planetary gear PG1 has a tooth ratio λ1 = 21 / 90 = 0.233.

[0126] The transmission mechanism FA0501 further includes brakes BF, B1, B2, B3, B4, and B5. Each of these brakes is a wet multi-plate type in which a rotating disc (the side to be braked) and a fixed disc are packed alternately, and the fixed disc is fitted into case K. Brake BF stops the rotation of ring gear RF of double pinion planetary gear PGF. Brake B1 stops the rotation of ring gear R1 of double pinion planetary gear PG1. The same is true for brakes B2, B3, B4, and B5.

[0127] <1st gear> As shown in FIG. 10B, in first gear (number 1 in a circle), brakes BF and B1 are engaged. Rotation from the input coupling mechanism IN enters the sun gear SF of the double pinion planetary gear PGF, and because the ring gear RF is stopped, the carrier CF rotates. The rotation of the carrier CF is input to the sun gear S1 of the double pinion planetary gear PG1. Because the ring gear R1 is stopped, the rotation of the sun gear S1 is transmitted to the carrier C1. The rotation of the carrier C1 is output to the output section G0.

[0128] For the double pinion planetary gear PGF, the ring gear rotation speed ωR = 0, so from the above formula (2) the carrier rotation speed ωC = -1.000ωS, and the reduction ratio JF = -1.000. For the double pinion planetary gear PG1, the ring gear rotation speed ωR = 0, so from the above formula (2) the carrier rotation speed ωC = -0.304ωS, and the reduction ratio J1 = -3.286. Therefore, the gear ratio i1 in 1st gear is as follows. i1=JF×J1=(-1.000)×(-3.286)≒3.29

[0129] <At reverse 1st gear> As shown in Fig. 10B, in first reverse gear (R1), the clutch hub sleeve CHS and the brake B1 are engaged. That is, the sleeve SL of the clutch hub sleeve CHS slides (indicated by a dotted line in Fig. 10A) and is engaged with the carrier CF of the double pinion planetary gear PGF. At this time, the carrier CF rotates in the same direction as the input IN. In other words, a direct connection state is established. The rotation of the carrier CF is input to the sun gear S1 of the double pinion planetary gear PG1. Because the ring gear R1 is stopped, the rotation of the sun gear S1 is transmitted to the carrier C1. The rotation of the carrier C1 is output to the output section G0.

[0130] For the double pinion planetary gear PG1, the reduction ratio J1 = -3.286 as described above. Therefore, the gear ratio iR1 in 1st reverse gear is as follows. iR1 = (1.000) × (-3.286) ≒ -3.29

[0131] In this way, the double pinion planetary gear PGF operates during forward travel. Since the double pinion planetary gear PGF has a tooth ratio λF=45 / 90, the reduction ratio is -1.000, and the input to the speed-changing planetary gear group (double pinion planetary gears PG1 to PG5) is a constant speed reverse rotation. On the other hand, during reverse travel, the carrier CF of the double pinion planetary gear PGF is directly connected to the input IN due to engagement of the clutch hub sleeve CHS, so that the input to the speed-changing planetary gear group (double pinion planetary gears PG1 to PG5) is a constant speed forward rotation. In this way, in this embodiment, the clutch hub sleeve CHS is used as a reverse gear select system.

[0132] In addition, for 2nd to 5th gears, the only difference is that double pinion planetary gears PG2 to PG5 operate instead of double pinion planetary gear PG1 that operates in 1st gear, so explanation will be omitted.Similarly, for reverse 2nd to 5th gears, the only difference is that double pinion planetary gears PG2 to PG5 operate instead of double pinion planetary gear PG1 that operates in reverse 1st gear, so explanation will be omitted.

[0133] <Effects> In this embodiment, the skeleton configuration as described above makes it possible to easily obtain the same number of gears for both forward and reverse.

[0134] In this embodiment, the clutch hub sleeve CHS is used for the reverse gear selection method, so that the reverse gear can be realized by a simple mechanism.

[0135] In this embodiment, the structure is simplified by braking the ring gear in each of the double pinion planetary gears PG1, PG2, PG3, PG4, PG5, and PGF. Since the ring gear is located on the outside close to the case, a configuration in which the ring gear is braked is advantageous in terms of layout.

[0136] <Other embodiments> Other embodiments will be briefly described below.

[0137] FIG. 11 shows a skeleton diagram and a brake application table of the transmission mechanism FA0601 according to another embodiment. The transmission mechanism FA0601 is a front-wheel drive 6-speed forward / 3-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 7 sets of double pinion planetary gears (DPPG). The reverse power flow is composed of 2 sets of DPPG, with the input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0138] FIG. 12 shows a skeleton diagram and a brake application table of the transmission mechanism FA0602 according to another embodiment. The transmission mechanism FA0602 is a front-wheel drive 6-speed forward / 3-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 7 DPPG sets. The reverse power flow is composed of 2 DPPG sets, and the input and output rotate in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0139] FIG. 13 shows a skeleton diagram and a brake application table of the transmission mechanism FA0604 according to another embodiment. The transmission mechanism FA0604 is a front-wheel drive 6-speed forward / 2-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 7 DPPG sets. The reverse power flow is composed of 2 DPPG sets, with the input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0140] FIG. 14 shows a skeleton diagram and a brake application table of the transmission mechanism FA0605 according to another embodiment. The transmission mechanism FA0605 is a 6 forward speed / 2 reverse speed transmission mechanism for front-wheel drive. The main gear train is arranged on one shaft. It is composed of 7 DPPG sets. The reverse power flow is composed of 1 DPPG set, and the input and output rotate in the same direction. Reverse 1st and 2nd speeds and forward 1st and 2nd speeds are additional reduction gears. The additional reduction DPPG rotates in the same direction as the input and output.

[0141] The additional reduction method involves additional reduction in speed because the gear ratio of the planetary gears for low-speed gear shifting alone does not achieve the desired gear ratio for an automatic transmission. For the FA0605 transmission, this corresponds to brake BZ in the brake application table. The reduction ratio at DPPG when brake BR is activated is 1.765, and the reduction ratio at DPPG when brake B12 is activated is -1.727, which is insufficient to achieve the reduction ratio required for an automatic transmission, so an additional reduction ratio of 2.000 is applied by activating brake BZ.

[0142] FIG. 15 shows a skeleton diagram and a brake application table of the transmission mechanism FA0606 according to another embodiment. The transmission mechanism FA0606 is a 6 forward speed / 2 reverse speed transmission mechanism for front-wheel drive. The main gear train is arranged on one shaft. It is configured by combining 6 sets of DPPG and CHS. CHS is applied to the reverse power flow. Reverse 1st and 2nd speeds and forward 1st and 2nd speeds are additional reduction gear systems. The additional reduction DPPG rotates in the same direction as the input and output.

[0143] 16 shows a skeleton diagram and a brake application table of the transmission mechanism FA0801 according to another embodiment. The transmission mechanism FA0801 is a front-wheel drive transmission mechanism with 8 forward speeds and 2 reverse speeds. The main gear train has a single shaft arrangement. It is configured by combining 5 sets of double pinion planetary gears (DPPG) and 2 sets of single pinion planetary gears (SPPG).

[0144] FIG. 17 shows a skeleton diagram and a brake application table of the transmission mechanism FA0802 according to another embodiment. The transmission mechanism FA0802 is a front-wheel drive 8-speed forward / 2-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 8 DPPG sets. The reverse power flow is composed of 2 DPPG sets, with the input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0145] FIG. 18 shows a skeleton diagram and a brake application table of the transmission mechanism FA0803 according to another embodiment. The transmission mechanism FA0803 is a front-wheel drive 8-speed forward / 4-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 8 DPPG sets. The reverse power flow is composed of 2 DPPG sets, with the input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0146] Fig. 19 shows a skeleton diagram and a brake application table of the transmission mechanism FA0804 according to another embodiment. The transmission mechanism FA0804 is a front-wheel drive 8-speed forward / 4-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 8 DPPG sets. The reverse power flow is composed of 2 DPPG sets, with the input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0147] Fig. 20 shows a skeleton diagram and a brake application table of the transmission mechanism FA0805 according to another embodiment. The transmission mechanism FA0805 is a front-wheel drive 8-speed forward / 4-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is configured with 9 DPPG sets. Reverse 1-4 and forward 1-4 speeds are additional reduction gears. The additional reduction DPPG rotates in the same direction as the input and output.

[0148] FIG. 21 shows a skeleton diagram and a brake application table of the transmission mechanism FA0806 according to another embodiment. The transmission mechanism FA0806 is a front-wheel drive 8-speed forward / 4-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is configured as a combination of 6 DPPG sets, 1 SPPG set, and CHS. The CHS is applied to the reverse power flow. Reverse 1-4 and forward 1-4 speeds use an additional reduction gear system. The additional reduction gear DPPG rotates in the same direction as the input and output. Note that when using a CHS for reverse, like this transmission mechanism FA0806, the additional reduction gear system is essential.

[0149] FIG. 22 shows a skeleton diagram and a brake application table of the transmission mechanism FA0901 according to another embodiment. The transmission mechanism FA0901 is a front-wheel drive 9-speed forward / 3-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 8 DPPG sets. The reverse power flow is composed of 2 DPPG sets, with the input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0150] FIG. 23 shows a skeleton diagram and brake application table of the transmission mechanism FA0902 according to another embodiment. The transmission mechanism FA0902 is a front-wheel drive 9-speed forward / 3-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 8 DPPG sets. The reverse power flow is composed of 2 DPPG sets, with the input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state. The output of the reverse DPPG and the output of the forward DPPG are individually linked to the AT output gear.

[0151] FIG. 24 shows a skeleton diagram and a brake application table of the transmission mechanism FA0903 according to another embodiment. The transmission mechanism FA0903 is a front-wheel drive 9-speed forward / 3-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is composed of 8 DPPG sets. The reverse power flow is composed of 2 DPPG sets, with the input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0152] 25 shows a skeleton diagram and a brake application table of the transmission mechanism FB0601 according to another embodiment. The transmission mechanism FB0601 is a front-wheel drive transmission mechanism with 6 forward speeds and 1 reverse speed. The main gear train is arranged on one shaft. It is a combination of 4 sets of DPPG and a Ravigneaux type planetary gear (Ravigneaux PG).

[0153] 26 shows a skeleton diagram and a brake application table of the transmission mechanism FB0602 according to another embodiment. The transmission mechanism FB0602 is a front-wheel drive transmission mechanism with 6 forward speeds and 1 reverse speed. The main gear train is arranged on one shaft. It is a combination of 3 sets of DPPG and Ravigneaux PG.

[0154] 27 shows a skeleton diagram and a brake application table of the transmission mechanism FB0603 according to another embodiment. The transmission mechanism FB0603 is a front-wheel drive transmission mechanism with 6 forward speeds and 1 reverse speed. The main gear train is arranged on one shaft. It is a combination of 3 sets of DPPG and Ravigneaux PG.

[0155] FIG. 28 shows a skeleton diagram and a brake application table of the transmission mechanism FB0604 according to another embodiment. The transmission mechanism FB0604 is a front-wheel drive 6-speed forward / 1-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is a combination of 3 sets of DPPG and Ravigneau PG. The DPPG for input to the sun gears S1 and S2 of the Ravigneau PG is an upsized version.

[0156] FIG. 29 shows a skeleton diagram and a brake application table of the transmission mechanism FB0801 according to another embodiment. The transmission mechanism FB0801 is a front-wheel drive 8-speed forward / 2-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is a combination of 4 sets of DPPG and Ravigneau PG. The inputs to the Ravigneau PG are S1×2 (2 inputs to sun gear S1), S2×2 (2 inputs to sun gear S2), and no input to C (carrier). The following descriptions are abbreviated in the same manner.

[0157] Fig. 30 shows a skeleton diagram and a brake application table of the transmission mechanism FB0802 according to another embodiment. The transmission mechanism FB0802 is a front-wheel drive transmission mechanism with 8 forward speeds and 1 reverse speed. The main gear train is arranged on one shaft. It is a combination of 4 sets of DPPG and Ravigneau PG. The inputs to the Ravigneau PG are S1 x 3, S2 x 1, and no input to C.

[0158] FIG. 31 shows a skeleton diagram and a brake application table of the transmission mechanism FB0901 according to another embodiment. The transmission mechanism FB0901 is a front-wheel drive 9-speed forward / 1-speed reverse transmission mechanism. The main gear train is arranged on one shaft. It is a combination of 4 sets of DPPG and Ravigneau PG. The inputs to the Ravigneau PG are S1×3, S2×1, and no input to C.

[0159] FIG. 32 shows a skeleton diagram and a brake application table of the transmission mechanism FB1001 according to another embodiment. The transmission mechanism FB1001 is a front-wheel drive 10 forward speed / 1 reverse speed transmission mechanism. The main gear train is arranged on one shaft. It is configured by combining 4 sets of DPPGs and a Ravigneaux PG. The inputs to the Ravigneaux PG are S1×2, S2×1, and C×1. One of the two DPPGs at the S1 input of the Ravigneaux PG is an upsized version.

[0160] FIG. 33 shows a skeleton diagram and a brake application table of the transmission mechanism FC0601 according to another embodiment. The transmission mechanism FC0601 is a front-wheel drive 6-speed forward / 2-speed reverse transmission mechanism. The main gear train is a 3-shaft arrangement (#1, #2, #3). It is a combination of 6 DPPG sets and CHS. It has 2-shaft output, and there are two types of final gear pinions, one for high speed (FGP-Hi) and one for low speed (FGP-Lo). A CHS is applied to the reverse power flow.

[0161] FIG. 34 shows a skeleton diagram and a brake application table of the transmission mechanism FC0602 according to another embodiment. The transmission mechanism FC0602 is a front-wheel drive 6-speed forward / 2-speed reverse transmission mechanism. The main gear train has a 3-shaft arrangement (#1, #2, #3). It is configured with 7 sets of DPPG. It has 2-shaft output, and there are two types of final gear pinions, one for high speed and one for low speed. The reverse power flow is configured with 2 sets of DPPG. When traveling forward, the reverse DPPG is in a completely free state.

[0162] FIG. 35 shows a skeleton diagram and brake application table of the transmission mechanism FC0603 according to another embodiment. The transmission mechanism FC0603 is a front-wheel drive transmission with 6 forward speeds and 2 reverse speeds. The main gear train has a 2-shaft arrangement (#1, #2). It is a combination of 4 DPPG sets and 2 SPPG sets. There are two types of transfer gear trains TG. For the inter-shaft rotation transmission between shafts #1 and #2, TG1 transmits in reverse and TG2 transmits in forward.

[0163] FIG. 36 shows a skeleton diagram and a brake application table of the transmission mechanism FC0604 according to another embodiment. The transmission mechanism FC0604 is a front-wheel drive 6-speed forward / 2-speed reverse transmission mechanism. The main gear train has a 2-shaft arrangement (#1, #2). It is configured with 7 DPPG sets. The reverse power flow is configured with 2 DPPG sets. When traveling forward, the reverse DPPG is in a completely free state.

[0164] FIG. 37 shows a skeleton diagram and a brake application table of the transmission mechanism FC0801 according to another embodiment. The transmission mechanism FC0801 is a front-wheel drive 8-speed forward / 4-speed reverse transmission mechanism. The main gear train has a 3-shaft arrangement (#1, #2, #3). It is configured with 8 sets of DPPG. It has 2-shaft output. The reverse power flow is configured with 2 sets of DPPG. When traveling forward, the reverse DPPG is in a completely free state.

[0165] FIG. 38 shows a skeleton diagram and brake application table of the transmission mechanism FC0802 according to another embodiment. The transmission mechanism FC0802 is a front-wheel drive 8-speed forward / 2-speed reverse transmission mechanism. The main gear train has a 3-shaft arrangement (#1, #2, #3). It is configured with 9 sets of DPPG. Fixed reduction is applied to reverse 1st and 2nd gears and forward 1st and 2nd gears. It has 2-shaft output, and there are two types of final gear pinions, one for high speed and one for low speed. The reverse power flow is configured with 2 sets of DPPG. When traveling forward, the reverse DPPG is in a completely free state.

[0166] FIG. 39 shows a skeleton diagram and a brake application table of the transmission mechanism FC0901 according to another embodiment. The transmission mechanism FC0901 is a front-wheel drive transmission mechanism with 9 forward speeds and 3 reverse speeds. The main gear train has a 3-shaft arrangement (#1, #2, #3). It is configured with 8 sets of DPPG. It has 2-shaft output, and there are two types of final gear pinions, one for high speed and one for low speed. The reverse power flow is configured with 2 sets of DPPG. When traveling forward, the reverse DPPG is in a completely free state.

[0167] FIG. 40 shows a skeleton diagram and a brake application table of the transmission mechanism FC0902 according to another embodiment. The transmission mechanism FC0902 is a front-wheel drive 9-speed forward / 3-speed reverse transmission mechanism. The main gear train has a 2-shaft arrangement (#1, #2). It is configured with 8 DPPG sets. The reverse power flow is configured with 2 DPPG sets. When traveling forward, the reverse DPPG is in a completely free state.

[0168] FIG. 41 shows a skeleton diagram and brake application table of the transmission mechanism FD0601 according to another embodiment. The transmission mechanism FD0601 is a front-wheel drive 6-speed forward / 1-speed reverse transmission mechanism. The main gear train is arranged on two shafts. It is a combination of 3 sets of DPPG and Ravigneaux PG. The transfer gear train TG has two types (TG1, TG2) of reduction and constant speed, which are arranged adjacent to each other.

[0169] FIG. 42 shows a skeleton diagram and brake application table of the transmission mechanism FD0801 according to another embodiment. The transmission mechanism FD0801 is a front-wheel drive transmission mechanism with 8 forward speeds and 2 reverse speeds. The main gear train has a two-shaft arrangement. It is a combination of 4 sets of DPPG and Ravigneau PG. The transfer gear train TG has a constant speed adjacent arrangement. The inputs to the Ravigneau PG are S1×2, S2×2, and no input to C.

[0170] Figure 43 shows a skeleton diagram and a brake application table of the transmission mechanism FD0901 according to another embodiment. The transmission mechanism FD0901 is a front-wheel drive transmission mechanism with 9 forward speeds and 3 reverse speeds. The main gear train has a two-shaft arrangement. It is a combination of 6 DPPG sets and a Ravigneaux PG.

[0171] FIG. 44 shows a skeleton diagram and a brake application table of the transmission mechanism FD0902 according to another embodiment. The transmission mechanism FD0902 is a front-wheel drive 9-speed forward and 2-speed reverse transmission mechanism. The main gear train has a two-shaft arrangement. It is a combination of 5 DPPG sets and a Ravigneaux PG. The inputs to the Ravigneaux PG are S1×2, S2×2, and C×1.

[0172] FIG. 45 shows a skeleton diagram and a brake application table of the transmission mechanism FD0903 according to another embodiment. The transmission mechanism FD0903 is a front-wheel drive transmission mechanism with 9 forward speeds and 1 reverse speed. The main gear train has a two-shaft arrangement. It is a combination of 4 sets of DPPG and Ravigneau PG. The inputs to the Ravigneau PG are S1×2, S2×1, and C×1.

[0173] Fig. 46 shows a skeleton diagram and a brake application table of the transmission mechanism RA0801 according to another embodiment. The transmission mechanism RA0801 is a rear-wheel drive transmission mechanism with 8 forward speeds and 4 reverse speeds. It is composed of 8 sets of DPPG. The reverse power flow is composed of 2 sets of DPPG, and the input and output rotate in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0174] 47 shows a skeleton diagram and a brake application table of a speed change mechanism RA0802 according to another embodiment. The speed change mechanism RA0802 is a rear-wheel drive 8-speed forward / 4-speed reverse speed change mechanism. It is composed of 6 DPPG sets and 1 SPPG set.

[0175] Fig. 48 shows a skeleton diagram and a brake application table of the transmission mechanism RA0803 according to another embodiment. The transmission mechanism RA0803 is a rear-wheel drive 8-speed forward / 4-speed reverse transmission mechanism. It is a combination of 6 DPPG sets, 1 SPPG set, and CHS. The SPPG has a fixed reduction gear function for reverse, and outputs reverse power flow when the CHS is activated.

[0176] Fig. 49 shows a skeleton diagram and a brake application table of the speed change mechanism RA0804 according to another embodiment. The speed change mechanism RA0804 is a rear-wheel drive 8-speed forward / 4-speed reverse speed change mechanism. It is a combination of 6 DPPG sets, 1 SPPG set, and CHS. The SPPG outputs 1st to 4th forward speeds and 1st to 4th reverse speeds at reduced speeds. When traveling in 5th to 8th forward speeds, the SPPG is completely free.

[0177] Figure 50 shows a skeleton diagram and a brake application table of the transmission mechanism RA0805 according to another embodiment. The transmission mechanism RA0805 is a rear-wheel drive transmission mechanism with 8 forward speeds and 2 reverse speeds. It is a combination of 7 sets of DPPG and CHS. The most upstream DPPG has a forward rotation fixed reduction function exclusively for reverse, and outputs reverse power flow when the CHS is activated.

[0178] Fig. 51 shows a skeleton diagram and a brake application table of the transmission mechanism RA0901 according to another embodiment. The transmission mechanism RA0901 is a rear-wheel drive transmission mechanism with 9 forward speeds and 3 reverse speeds. It is composed of 8 DPPG sets. The reverse power flow is composed of 2 DPPG sets, and the input and output rotate in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0179] 52 shows a skeleton diagram and a brake application table of a speed change mechanism RA0902 according to another embodiment. The speed change mechanism RA0902 is a rear-wheel drive speed change mechanism with 9 forward speeds and 3 reverse speeds. It is composed of 6 DPPG sets and 1 SPPG set.

[0180] 53 shows a skeleton diagram and a brake application table of a speed change mechanism RA0903 according to another embodiment. The speed change mechanism RA0903 is a rear-wheel drive speed change mechanism with 9 forward speeds and 3 reverse speeds. It is composed of 6 DPPG sets and 1 SPPG set.

[0181] 54 shows a skeleton diagram and a brake application table of the transmission mechanism RA1001 according to another embodiment. The transmission mechanism RA1001 is a rear-wheel drive 10-speed forward / 5-speed reverse transmission mechanism. It is a combination of 7 DPPG sets, 1 SPPG set, and CHS. The SPPG has a fixed reduction gear function for reverse, and outputs reverse power flow when the CHS is activated.

[0182] Fig. 55 shows a skeleton diagram and a brake application table of the transmission mechanism RA1201 according to another embodiment. The transmission mechanism RA1201 is a rear-wheel drive transmission mechanism with 12 forward speeds and 3 reverse speeds. It is composed of 9 DPPG sets. The reverse power flow is composed of 2 DPPG sets, and the input and output rotate in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0183] Fig. 56 shows a skeleton diagram and a brake application table of the transmission mechanism RB0601 according to another embodiment. The transmission mechanism RB0601 is a rear-wheel drive transmission mechanism with 6 forward speeds and 1 reverse speed. It is a combination of 4 sets of DPPGs and a Ravigneaux PG. The DPPG at the most upstream inputs the reverse deceleration output to the transmission function unit.

[0184] FIG. 57 shows a skeleton diagram and a brake application table of the transmission mechanism RB0801 according to another embodiment. The transmission mechanism RB0801 is a rear-wheel drive transmission mechanism with 8 forward speeds and 1 reverse speed. It is a combination of 5 DPPG sets and a Ravigneaux PG. The DPPG at the most upstream inputs the reverse deceleration output to the transmission function unit. The inputs to the Ravigneaux PG are S1×2, S2×1, and C×1.

[0185] Fig. 58 shows a skeleton diagram and brake application table of the transmission mechanism RB0802 according to another embodiment. The transmission mechanism RB0802 is a rear-wheel drive transmission mechanism with 8 forward speeds and 1 reverse speed. It is a combination of 5 DPPG sets and a Ravigneaux PG. The DPPG at the most upstream inputs the reverse deceleration output to the transmission function unit. The inputs to the Ravigneaux PG are S1 x 3, S2 x 1, and no input to C.

[0186] FIG. 59 shows a skeleton diagram and brake application table of the transmission mechanism RB0901 according to another embodiment. The transmission mechanism RB0901 is a rear-wheel drive 9-speed forward / 1-reverse speed transmission mechanism. It is a combination of 5 DPPG sets and a composite PG (Simpson type). The most upstream DPPG inputs the reverse deceleration output to the transmission function unit. The inputs to the composite PG are 2 sun gears and 1 other.

[0187] Figure 60 shows a skeleton diagram and brake application table of the transmission mechanism RB0902 according to another embodiment. The transmission mechanism RB0902 is a rear-wheel drive transmission mechanism with 9 forward speeds and 2 reverse speeds. It is a combination of 6 DPPG sets and a Ravigneaux type PG. The upstream DPPG inputs the reverse deceleration output to the transmission function unit. The inputs to the Ravigneaux type PG are S1 x 2, S2 x 2, and C x 1.

[0188] FIG. 61 shows a skeleton diagram and a brake application table of the transmission mechanism RB0903 according to another embodiment. The transmission mechanism RB0903 is a rear-wheel drive transmission mechanism with 9 forward speeds and 2 reverse speeds. It is a combination of 5 DPPG sets and a Ravigneaux type PG. The two upstream DPPGs input reverse deceleration outputs to the transmission function section. The inputs to the Ravigneaux type PG are S1×2, S2×2, and C×2.

[0189] FIG. 62 shows a skeleton diagram and a brake application table of the transmission mechanism RB0904 according to another embodiment. The transmission mechanism RB0904 is a rear-wheel drive 9-speed forward / 3-speed reverse transmission mechanism. It is a combination of 6 DPPG sets and a Ravigneaux type PG. The three upstream DPPGs input reverse deceleration and constant speed output to the transmission function section. The inputs to the Ravigneaux type PG are S1×3, S2×3, and C×3.

[0190] Figure 63 shows a skeleton diagram and brake application table of the transmission mechanism RB0905 according to another embodiment. The transmission mechanism RB0905 is a rear-wheel drive transmission mechanism with 9 forward speeds and 3 reverse speeds. It is a combination of 6 DPPG sets and a Ravigneaux type PG. The inputs to the Ravigneaux type PG are S1 x 4, S2 x 4, and no input to C.

[0191] Fig. 64 shows a skeleton diagram and a brake application table of the transmission mechanism RB1001 according to another embodiment. The transmission mechanism RB1001 is a rear-wheel drive transmission mechanism with 10 forward speeds and 1 reverse speed. It is a combination of 5 DPPG sets and a Ravigneaux type PG. The inputs to the Ravigneaux type PG are S1 x 2, S2 x 1, and C x 1.

[0192] Figure 65 shows a skeleton diagram and brake application table of a transmission mechanism RB1002 according to another embodiment. The transmission mechanism RB1002 is a rear-wheel drive transmission mechanism with 10 forward speeds and 1 reverse speed. It is configured by combining five DPPG sets with a Ravigneaux type PG. The inputs to the Ravigneaux type PG are S1 x 2, S2 x 1, and C x 1. The rearmost DPPG reverses and decelerates the ring gear output (all gear positions) of the Ravigneaux PG to produce the output of the transmission.

[0193] FIG. 66 shows a skeleton diagram and brake application table of a transmission mechanism RB1003 according to another embodiment. The transmission mechanism RB1003 is a rear-wheel drive transmission mechanism with 10 forward speeds and 2 reverse speeds. It is configured by combining 5 DPPG sets and a Ravigneaux type PG. The two upstream DPPGs input reverse deceleration and acceleration outputs to the transmission function unit. The inputs from the transmission function unit to the Ravigneaux type PG are S1×2, S2×1, and no input to C.

[0194] FIG. 67 shows a skeleton diagram and a brake application table of the transmission mechanism RB1201 according to another embodiment. The transmission mechanism RB1201 is a 12-speed forward / 2-speed reverse transmission mechanism for rear-wheel drive. It is a combination of 5 DPPG sets and a Ravigneaux type PG. The inputs to the Ravigneaux type PG in the transmission function section are S1×1, S2×1, and C×1. The two upstream DPPGs input reverse deceleration outputs to the transmission function section.

[0195] Figure 68 shows a skeleton diagram and brake application table of the transmission mechanism RB1202 according to another embodiment. The transmission mechanism RB1202 is a 12-speed forward / 2-speed reverse transmission mechanism for rear-wheel drive. It is configured by combining 6 DPPG sets with a Ravigneaux type PG. The inputs to the Ravigneaux type PG from the four DPPGs in the transmission function section are S1 x 3, S2 x 1, and no input to C. The two upstream DPPGs input reverse deceleration and acceleration outputs to the transmission function section.

[0196] The speed change mechanism of the stepped automatic transmission for an automobile according to the present invention is not limited to the above-described embodiment, and various modifications and improvements are possible within the scope of the claims. [Explanation of symbols]

[0197] PG11,PG12,PG21,PG22,PGR1,PGR2,PG13,PG46,PGA,PGB,PGGC,PGZ,PGF,PG1,PG2,PG3,PG4,PG5...Double pinion planetary gear,PG0...Lavigneaux planetary gear, B11, B12, B21, B22, B01, B02, BR1, BR2, B13, B46, BA, BB, BC, BF, B1, B2, B3, B4, B5...Brakes, Z···Ring gear fixing member, CHS···Clutch hub sleeve, G1: First planetary gear group, G2: Second planetary gear group, G3: Third planetary gear group

Claims

1. A gear train is formed by connecting a plurality of planetary gears, Each of the plurality of planetary gears includes a sun gear, a ring gear, and a carrier that rotatably supports a plurality of pinion gears, Each of the planetary gears is operated only by a brake without using a clutch, the plurality of planetary gears include a compound planetary gear, 2. A transmission mechanism for a stepped automatic transmission for an automobile, comprising: a first input member for inputting a plurality of speed change speeds to one input member of said compound planetary gear;

2. A gear train is formed by connecting a plurality of planetary gears, Each of the plurality of planetary gears includes a sun gear, a ring gear, and a carrier that rotatably supports a plurality of pinion gears, Each of the planetary gears is operated only by a brake without using a clutch, A transmission mechanism for a stepped automatic transmission for an automobile, characterized in that it has a plurality of reverse gears.

3. A gear train is formed by connecting a plurality of planetary gears, Each of the plurality of planetary gears includes a sun gear, a ring gear, and a carrier that rotatably supports a plurality of pinion gears, Each of the planetary gears is operated only by a brake without using a clutch, the plurality of planetary gears include two single-row double pinion planetary gears connected in series, The power flow of the reverse gear is formed by the two single-row double pinion planetary gears, a gear shift mechanism for a stepped automatic transmission for an automobile, characterized in that the input rotation direction and the output rotation direction of said two single-row double pinion planetary gears are the same when viewed as a group.

4. A gear train is formed by connecting a plurality of planetary gears, Each of the plurality of planetary gears includes a sun gear, a ring gear, and a carrier that rotatably supports a plurality of pinion gears, Each of the planetary gears is operated only by a brake without using a clutch, Further comprising a clutch hub sleeve, A transmission mechanism for a stepped automatic transmission for an automobile, characterized in that the above clutch hub sleeve is used as a reverse gear selection system.

5. A gear train is formed by connecting a plurality of planetary gears, Each of the plurality of planetary gears includes a sun gear, a ring gear, and a carrier that rotatably supports a plurality of pinion gears, Each of the planetary gears is operated only by a brake without using a clutch, The plurality of planetary gears include two planetary gear groups each having a reverse rotation output function and each being composed of a plurality of single-row planetary gears, A transmission mechanism for a stepped automatic transmission for an automobile, characterized in that two of the planetary gear groups are connected in series to provide a normal rotation output as a transmission.

6. A gear train is formed by connecting a plurality of planetary gears, Each of the plurality of planetary gears includes a sun gear, a ring gear, and a carrier that rotatably supports a plurality of pinion gears, Each of the planetary gears is operated only by a brake without using a clutch, the plurality of planetary gears include a first planetary gear group consisting of a plurality of single-row planetary gears and having a reverse rotation output function, and a second planetary gear group consisting of another plurality of single-row planetary gears and having a forward rotation output function, a gear shift mechanism for a stepped automatic transmission for an automobile, the gear shift mechanism being characterized in that the first planetary gear group and the second planetary gear group are connected in series to provide a reverse rotation output as a transmission.

7. A gear train is formed by connecting a plurality of planetary gears, Each of the plurality of planetary gears includes a sun gear, a ring gear, and a carrier that rotatably supports a plurality of pinion gears, Each of the planetary gears is operated only by a brake without using a clutch, the plurality of planetary gears includes a plurality of single-row planetary gears, a brake for controlling each of the plurality of single-row planetary gears brakes the ring gear of each of the single-row planetary gears,

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