Transmission mechanism for stepped automatic transmissions in automobiles

Abstract

To provide a gear shifting mechanism for a stepped automatic transmission for automobiles that enables structural simplification, miniaturization, weight reduction, and cost reduction by eliminating the use of a clutch. [Solution] Multiple planetary gears (PG11, PG12, PG21, PG0) are linked together to form a gear train. Each of the multiple 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 multiple pinion gears. Each of the multiple planetary gears is operated solely by brakes (B11, B12, B21, B01, B02) without the use of a clutch.

Description

Technical Field

[0001] The present invention relates to a transmission mechanism of a stepped automatic transmission for automobiles.

Background Art

[0002] As a transmission mechanism of a stepped automatic transmission for automobiles and the like, a plurality of double pinion planetary gear sets are provided, and as friction elements for controlling engagement and release during shifting, a clutch interposed between rotating elements to engage the rotating elements with each other and a brake interposed between a rotating element and a case to stop the rotating element are provided (see, for example, Patent Document 1). In addition, a transmission using an incomplete planetary gear mechanism having two parts, one central gear and one planetary carrier, has also been proposed (see, for example, Patent Document 2).

Prior Art Documents

Patent Documents

[0003]

Patent Document 1

Patent Document 2

Summary of the Invention

Problems to be Solved by the Invention

[0004] Generally, a wet multi-plate clutch is used as the clutch. However, the mechanism of the clutch has the following problems. That is, the rotating clutch drum is heavy, a mechanism for canceling the residual pressure at the time of release is required, the support mechanism of the clutch is complicated, heavy, and disadvantageous in terms of space, the oil passage structure for supplying pressure oil to the rotating body is complicated, a seal ring for preventing pressure oil leakage is required, and the degree of freedom in the layout of the entire automatic transmission is reduced.

[0005] Therefore, the present invention aims to provide a gear shifting mechanism for a stepped automatic transmission for automobiles that enables simplification of structure, miniaturization, weight reduction, and cost reduction by eliminating the use of a clutch. [Means for solving the problem]

[0006] To achieve the above objective, the present invention provides a gear shift mechanism for a stepped automatic transmission for automobiles, in which a gear train is constructed by linking multiple planetary gears, each of the multiple planetary gears comprises a sun gear, a ring gear, and a carrier that rotatably supports multiple pinion gears, and each of the multiple planetary gears is operated solely by brakes without the use of a clutch. Here, the gear shift mechanism of the present invention is intended for stepped automatic transmissions for automobiles powered by an engine. The gear shift mechanism of the present invention comprises multiple planetary gears, but the individual planetary gears are common, so there are no problems with processing and assembly, and it is highly reliable. Multiple planetary gears are linked together and function as a multi-stage transmission by selecting and combining the brakes to be applied. Furthermore, each planetary gear can be arranged in close contact, and does not require space such as when using a pinion shaft gear that has multiple gears and integrates them (see Patent Document 2 above). Existing control function structures for stepped automatic transmissions in automobiles either use only the clutch or a combination of the clutch and brake; there are no systems that use only the brake. The inventors of this invention believe that this invention is unique in the structure of stepped automatic transmissions for automobiles.

[0007] Here, the plurality of planetary gears include a composite planetary gear, and it is preferable to input multiple gear speeds to one input member of the composite planetary gear. More specifically, the first output member of the first planetary gear and the second output member of the second planetary gear are connected to one input member of the composite planetary gear. The input from the first planetary gear and the input from the second planetary gear are switched to be input to the one input member by the operation of either the first brake provided on the first planetary gear or the second brake provided on the second planetary gear.

[0008] Furthermore, it is preferable to have multiple reverse gear stages.

[0009] Furthermore, the plurality of planetary gears preferably include two single-row double-pinion planetary gears connected in series, with the power flow of the reversing gear being composed of the two single-row double-pinion planetary gears, and the input rotation direction and output rotation direction when the two single-row double-pinion planetary gears are viewed as a group being the same.

[0010] Furthermore, it is preferable to further include a clutch hub sleeve and use the clutch hub sleeve as a reverse gear selector.

[0011] Furthermore, the plurality of planetary gears preferably consist of a group of planetary gears comprising a plurality of single-row planetary gears, including two planetary gear groups having a reverse output function, and the two planetary gear groups are connected in series to provide forward rotation output as a transmission.

[0012] Furthermore, the plurality of planetary gears includes a first planetary gear group consisting of a plurality of single-row planetary gears having a reverse output function, and a second planetary gear group consisting of other plurality of single-row planetary gears having a forward output function, and it is preferable to connect the first planetary gear group and the second planetary gear group in series to provide a reverse output as a transmission.

[0013] Furthermore, the plurality of planetary gears includes a plurality of single-row planetary gears, and for all of the plurality of single-row planetary gears, it is preferable that the brake that controls each single-row planetary gear brakes the ring gear of each single-row planetary gear. [Effects of the Invention]

[0014] According to the gear shifting mechanism for a stepped automatic transmission for automobiles described in claim 1, since each of the multiple planetary gears is operated solely by brakes without using a clutch, it is possible to simplify the structure, make it smaller, lighter, and reduce costs. In addition, the overall layout flexibility of the automatic transmission is improved.

[0015] According to the gear shift mechanism for a stepped automatic transmission for automobiles described in claim 2, since multiple gear speeds are input to one input member of a composite planetary gear, the gear ratio can be increased with a simple structure.

[0016] The Ravigno planetary gear, a type of composite planetary gear, is a compact composite 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 used as 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 gear speeds (for example, two) are input to the small sun gear S1, it is possible to obtain 2 x 2 = 4 reduction ratios.

[0017] The transmission mechanism for a stepped automatic transmission for automobiles described in claim 3 has multiple reverse gears. Generally, there is often only one reverse gear, in which case the gear ratio is similar to that of the first forward gear, resulting in a large driving force. This raises safety concerns when reversing or parking in a garage, but by having multiple reverse gears, it is possible to use a gear with appropriate driving force even when reversing. In this invention, since only brakes are used and no clutch is used, the overall structure of the automatic transmission is compact and has a high degree of freedom in layout, making it easy to realize multiple reverse gears from a layout perspective.

[0018] According to the gear shifting mechanism for a stepped automatic transmission for automobiles described in claim 4, the power flow of the reverse gear is configured by two single-row double pinion planetary gears connected in series. In each single-row double pinion planetary gear, if one of the sun gear and carrier is the input and the other is the output, the input rotation direction and the output rotation direction are opposite. 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 gear shifting mechanism for a stepped automatic transmission for automobiles described in claim 5, a clutch hub sleeve is used as the reverse gear selector. Therefore, a reverse gear can be realized with a simple mechanism.

[0020] According to the gear shifting mechanism for a stepped automatic transmission for an automobile described in claim 6, two planetary gear groups, each consisting of multiple single-row planetary gears and having a reverse output function, are connected in series to form a gear shift with forward rotation output. This allows, for example, output rotation in the same direction as the input rotation direction to be obtained when moving forward.

[0021] According to the transmission mechanism of the stepped automatic transmission for automobiles described in claim 7, a first planetary gear group composed of a plurality of single-row planetary gears and having a reverse output function, and a second planetary gear group composed of other plurality of single-row planetary gears and having a forward output function are connected in series to form a transmission with a reverse output. As a result, for example, when reversing, an output rotation in the opposite direction to the input rotation direction can be obtained.

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

Brief Description of the Drawings

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

[0024] Before describing embodiments of the present invention, general information will be provided to facilitate understanding. <Velocity relationship formula for single-row single-pinion planetary gear> Generally, in the single-row single-pinion planetary gear shown in Figure 1, the number of teeth on the sun gear is ZS, the number of teeth on the ring gear is ZR, the rotational speed of the sun gear is ωS, the rotational speed of the carrier is ωC, and the rotational speed of the ring gear is ωR. Also, λ is the ratio of the number of teeth on the sun gear (λ = ZS / ZR). The speed relationship of the single-row single-pinion planetary gear is expressed by the following equation. Note that here, it is referred to as "single-row type" in contrast to the composite type planetary gear described later. (1+λ)ωC =λωS+ωR (Equation 1)

[0025] <Velocity relationship formula for single-row double-pinion planetary gear> Generally, in the single-row double-pinion planetary gear shown in Figure 2, the number of teeth on the sun gear is ZS, the number of teeth on the ring gear is ZR, the rotational speed of the sun gear is ωS, the rotational speed of the carrier is ωC, and the rotational speed of the ring gear is ωR. Let λ be the ratio of the number of teeth on the sun gear (λ = ZS / ZR). The speed relationship for the single-row double-pinion planetary gear is expressed by the following equation. (1-λ)ωC = -λωS+ωR (Equation 2)

[0026] <Velocity relationship formula for Ravigno-type planetary gears> Generally, in the Ravigno-type planetary gear shown in Figure 3, if the number of teeth of sun gear S1 is ZS1, the rotational speed of sun gear S1 is ωS1, the number of teeth of sun gear S2 is ZS2, the rotational speed of sun gear S2 is ωS2, the number of teeth of ring gear is ZR, the rotational speed of ring gear is ωR, the rotational speed of carrier is ωC, the tooth ratio of sun gear S1 is α = ZS1 / ZR, and the tooth ratio of sun gear S2 is β = ZS2 / ZR, then the following speed relationship equations hold for the Ravigno-type planetary gear. Note that the inner pinion that meshes with 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 Ravigno-type planetary gear is an example of a composite planetary gear. For column S1 (1-α)·ωC = -α·ωS1 + ωR (Equation 3A) For column S2: (1+β)·ωC = β·ωS2 + ωR (Equation 3B)

[0027] Furthermore, if ωi is the input rotational speed to the AT (automatic transmission) main unit via the coupling, ωo is the output rotational speed to the AT main unit (final pinion rotational speed), and i is the gear ratio as an AT (i = ωi / ωo), then the output reduction ratio (= ωi / ωR) when the input to the Ravigno-type planetary gear is the reduction input and the ring gear is the output is as shown in the equations in the table in Figure 4. Here, J1 is the input reduction ratio to sun gear S1, J2 is the input reduction ratio to sun gear S2, and J3 is the input reduction ratio to the carrier. Note that the equations in Figure 4 are well known equations.

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

[0029] <Clutch hub sleeve> This section describes the clutch hub sleeve (hereinafter also called CHS) mechanism. CHS is generally applied to manual transmissions. Figure 5B is a simplified cross-sectional view of the CHS. As shown in Figure 5B, the CHS comprises a hub H and a sleeve SL. The rotating body A, such as a gear, has a meshing portion AK with a spline groove on its outer circumference. The hub H has a meshing portion HK with a spline groove on its outer circumference. The sleeve SL has a meshing portion SK with a spline groove on its inner circumference. The sleeve SL is movable axially relative to the rotating body A and the hub H. In the state shown in the right diagram of Figure 5B, the sleeve SL is meshed only with the hub H. In this state, the rotating body A and the hub H can rotate independently. When the sleeve SL moves axially to the state shown in the left diagram of Figure 5B, the sleeve SL is meshed with both the rotating body A and the hub H. In this state, the rotating body A and the hub H rotate together as a single unit in the same direction and at the same rotational speed.

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

[0031] <First Embodiment> The FB0605 transmission mechanism for a stepped automatic transmission for automobiles according to the first embodiment of the present invention will be described with reference to Figures 7A and 7B. Figure 7A is a skeleton diagram of the FB0605 transmission mechanism for a 6-speed forward and 1-speed reverse automatic transmission for front-wheel drive. For simplicity, only one side (upper half) of the rotating body is shown with the rotation axis X as the center.

[0032] The FB0605 transmission mechanism comprises three sets of double pinion planetary gears PG11, PG12, PG21 and a Ravigno planetary gear PG0. The main gear train, consisting of the double pinion planetary gears PG11, PG12, PG21 and the Ravigno planetary gear PG0, is arranged on a single axis (rotation axis X). The double pinion planetary gears PG11, PG12, PG21 and the Ravigno planetary gear PG0 each have known configurations. The double pinion planetary gear PG11 comprises a sun gear S11, a ring gear R11, and a carrier C11. The double pinion planetary gear PG12 comprises a sun gear S12, a ring gear R12, and a carrier C12. Carriers C11 and C12 are connected. The double pinion planetary gear PG21 comprises a sun gear S21, a ring gear R21, and a carrier C21. The Ravigno planetary gear PG0 comprises two sun gears S1 and S2, a ring gear R0, and a carrier C0. Sun gear S1 is connected to carrier C11. Sun gear S2 is connected to carrier C21.

[0033] The tooth ratio λ11 of the double pinion planetary gear PG11 is 36 / 90 = 0.400. Here, the number "36" displayed on the sun gear S11 in Figure 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 applies to the numbers displayed on the other gears. The tooth ratio λ12 of the double pinion planetary gear PG12 is 45 / 90 = 0.500. The tooth ratio λ21 of the double pinion planetary gear PG21 is 45 / 90 = 0.500.

[0034] The FB0605 transmission mechanism is further equipped with brakes B11, B12, B21, B01, and B02. Each of these brakes is a wet multi-plate type, with rotating discs (the side being braked) and stationary discs packed alternately, and the stationary discs are fitted into case K. Brake B11 stops the rotation of the ring gear R11 of the double pinion planetary gear PG11. Brake B12 stops the rotation of the ring gear R12 of the double pinion planetary gear PG12. Brake B21 stops the rotation of the ring gear R21 of the double pinion planetary gear PG21. Brake B01 stops the rotation of the carrier C0 of the Ravigno planetary gear PG0. Brake B02 stops the rotation of the carrier C21 of the double pinion planetary gear PG21 and the sun gear S2 of the Ravigno planetary gear PG0.

[0035] The input coupling mechanism IN is a fluid coupling, torque converter, wet clutch, etc. 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. In addition, the sun gear S1 of the Ravigno 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 Ravigno planetary gear PG0. The sun gear S2 of the Ravigno 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 Ravigno planetary gear PG0. Furthermore, there is no input from the Ravigno Planetary Gear PG0 to the carrier C0.

[0036] The transmission mechanism FB0605 is connected to the output section G0. The output section G0 comprises the transmission mechanism output gear G1, output idler gear G2, and output gear set G3. The output gear set G3 drives a differential (not shown). The differential (not shown) includes a final reduction gear. In the figure, T indicates the rearmost part of the gear train that determines the gear ratio (transmission ratio). Hereafter, this will also be referred to as the T section or T-axis. The gear ratio at the T section is shown in the rightmost column of Figure 7B (the same applies to Figures 8B, 9B, and 10B described later).

[0037] The rotation of the ring gear R0 of the Ravigno planetary gear PG0 is transmitted to the output gear G1 of the transmission mechanism, and this rotation is then transmitted sequentially to the output idler gear G2, the output gear set G3, and the differential (not shown). Furthermore, in the brake application table in Figure 7B, for example, in the column for Brake B01, there is a circle (〇) in the columns for 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 (〇) in the columns for 3rd gear (circled number 3) to 6th gear (circled number 6), indicating that the brake is not applied (non-braking state).

[0038] <When reversing> As shown in Figure 7B, in reverse (R), brakes B01 and B21 are engaged. Since brakes B11 and B12 are not engaged, the double pinion planetary gears PG11 and PG12 do not operate, and rotation from the input coupling mechanism IN enters the sun gear S21 of the double pinion planetary gear PG21. Since 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 Ravigno planetary gear PG0. Since 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 G0.

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

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

[0041] <In 1st gear> As shown in Figure 7B, in 1st gear (circled number 1), brakes B01 and B11 are engaged. The rotation from the input coupling mechanism IN enters the sun gear S11 of the double pinion planetary gear PG11, and since 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 Ravigno planetary gear PG0. Since 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 the double pinion planetary gear PG11, since the ring gear rotation speed ωR = 0, from equation (2) above, the carrier rotation speed ωC = -0.667ωS, and the reduction ratio J11 = -1.500, which is input to the sun gear S1 of the Ravigno planetary gear PG0. The carrier C0 of the Ravigno planetary gear PG0 is fixed (ωC = 0).

[0043] The gear ratio i1 in 1st gear is calculated as follows: Rotational speed of sun gear S1: ωS1 = (1 / J11)ωi = (ωi / -1.500) = -0.667ωi The rotational speed of the sun gear S2 is: ωS2 = -(α / β)·(1 / J11)ωi = -(0.267 / 0.333)·(1 / -1.500)ωi=0.535ωi Output rotational speed (= rotational speed of ring gear R0): 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] <In 2nd gear> As shown in Figure 7B, in second gear (circled number 2), brakes B01 and B12 are engaged. The rotation from the input coupling mechanism IN enters the sun gear S12 of the double pinion planetary gear PG12, and since the ring gear R12 is stopped, the carrier C12 rotates. Since carrier C12 and carrier C11 are connected, the rotation of carrier C12 is input to the sun gear S1 of the Ravigno planetary gear PG0. Since carrier C0 is stopped, the rotation of sun gear S1 is transmitted to the ring gear R0. The rotation of ring gear R0 is output to the output section G0.

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

[0046] The gear ratio i2 in 2nd gear is calculated as follows: Rotational speed of sun gear S1: ωS1 = (1 / J12)ωi = (ωi / -1.000) = -1.000ωi The rotational speed of the sun gear S2 is: ωS2 = -(α / β)·(1 / J12)ωi =-(0.267 / 0.333)·(1 / -1.000)ωi=0.802ωi Output rotational speed (= rotational speed of ring gear R0): ωo = ωR = (α / J12)·ωi = (0.267 / -1.000) ωi = -0.267ωi (See Figure 4(a)) Gear ratio in 2nd gear: i2 = (ωi / ωo) = (ωi / -0.267ωi) ≈ -3.75

[0047] <In 3rd gear> As shown in Figure 7B, in 3rd gear (circled number 3), brakes B02 and B11 are engaged. The rotation from the input coupling mechanism IN enters the sun gear S11 of the double pinion planetary gear PG11, and since 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 Ravigno planetary gear PG0. The rotation of the sun gear S1 is transmitted from the short pinion to the long pinion, and since 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 the double pinion planetary gear PG11, the ring gear rotation speed ωR = 0, so from equation (2) above, the carrier rotation speed ωC = -0.667ωS, and the reduction ratio J11 = -1.500, which is input to the sun gear S1 of the Ravigno planetary gear PG0. The sun gear S2 of the Ravigno planetary gear PG0 is fixed (ωS2 = 0).

[0049] The gear ratio i3 in 3rd gear is calculated as follows: Rotational speed of sun gear S1: 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 rotational speed (= rotational speed of ring gear R0): 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) Gear ratio in 3rd gear: i3 = (ωi / ωo) = (ωi / -0.396ωi) ≈ -2.53

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

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

[0052] The gear ratio i4 in 4th gear is calculated as follows: Rotational speed of sun gear S1: 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 rotational speed (= rotational speed of ring gear R0): ωo=ωR={α(1+β) / (α+β)}·(1 / J12)ωi =(0.356 / 0.600)·(1 / -1.000)ωi=-0.593ωi (See Figure 4(b)) Gear ratio in 4th gear: i4 = (ωi / ωo) = (ωi / -0.593ωi) ≈ -1.69

[0053] <At 5th gear> As shown in Figure 7B, in 5th gear (circled number 5), 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 since 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 Ravigno planetary gear PG0. The reduction ratio J11 of the double pinion planetary gear PG11 is -1.500. Also, the rotation from the input coupling mechanism IN enters the sun gear S21 of the double pinion planetary gear PG21, and since 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 Ravigno planetary gear PG0. The reduction ratio J21 of the double pinion planetary gear PG21 is -1.000. In other words, the Ravigno Planetary Gear PG0 has two input systems, with rotation being input to the sun gears S1 and S2. 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 unit G0.

[0054] The gear ratio i5 in 5th gear is calculated as follows: Rotational speed of sun gear S1: ωS1 = (1 / J11)ωi = (ωi / -1.500) = -0.667ωi Rotational speed of sun gear S2: ωS2 = (1 / J21)ωi = (ωi / -1.000) = -1.000ωi Carrier C0 rotational speed: 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 rotational speed (= rotational speed of ring gear R0): 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)) Gear ratio in 5th gear: i5 = (ωi / ωo) = (ωi / -0.802ωi) ≈ -1.25

[0055] <At 6th gear> As shown in Figure 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 since the ring gear R12 is stopped, the carrier C12 rotates. The rotation of carrier C12 is input to the sun gear S1 of the Ravigno planetary gear PG0 via carrier C11. The reduction ratio J12 of the double pinion planetary gear PG12 is -1.000. Also, the rotation from the input coupling mechanism IN enters the sun gear S21 of the double pinion planetary gear PG21, and since the ring gear R21 is stopped, the carrier C21 rotates. The rotation of carrier C21 is input to the sun gear S2 of the Ravigno planetary gear PG0. The reduction ratio J21 of the double pinion planetary gear PG21 is -1.000. In other words, the Ravigno Planetary Gear PG0 has two input systems, with rotation being input to the sun gears S1 and S2. 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 unit G0.

[0056] The gear ratio i6 in 6th gear is calculated as follows: Rotational speed of sun gear S1: ωS1 = (1 / J12)ωi = (ωi / -1.000) = -1.000ωi Rotational speed of sun gear S2: ωS2 = (1 / J21)ωi = (ωi / -1.000) = -1.000ωi Carrier C0 rotational speed: 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 rotational speed (= rotational speed of ring gear R0): 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)) Gear ratio in 6th gear: i6 = (ωi / ωo) = (ωi / -1.000ωi) = -1.00

[0057] In this embodiment, the operation of each planetary gear PG11, PG12, PG21, and PG0 is performed solely by brakes, without the use of clutches. This simplifies the structure, allows for miniaturization, weight reduction, and cost reduction, and improves the overall layout flexibility of the automatic transmission. Furthermore, in each gear position, each brake is either in a braking state or a non-braking state according to the brake application table, resulting in simpler control.

[0058] Furthermore, in this embodiment, the structure is simplified by braking the ring gear in each double pinion planetary gear B11, B12, and B21. Since the ring gear is located on the outside, close to the case, the configuration that brakes the ring gear is advantageous in terms of layout.

[0059] Furthermore, in this embodiment, two gear speeds are input to the sun gear S1 of the Ravigno planetary gear PG0 (two-system input). Specifically, 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 Ravigno 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 brakes B11 and B12. In other words, since multiple gear speeds are input to one input member of the composite planetary gear, the gear ratio can be increased with a simple structure.

[0060] <Second Embodiment> The transmission mechanism FA0603 for a stepped automatic transmission for automobiles according to a second embodiment of the present invention will be described with reference to Figures 8A and 8B. Figure 8A is a skeleton diagram of the transmission mechanism FA0603 for a 6-speed forward and 3-speed reverse automatic transmission for front-wheel drive. For simplicity, only one side (upper half) of the rotating body is shown with the rotation axis X as the center.

[0061] The FA0603 transmission mechanism comprises seven sets of 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 Figure 8A, the main gear train, consisting of the double pinion planetary gears PGR1, PGR2, PG13, PG46, PGA, PGB, and PGC, is arranged on a single axis.

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

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

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

[0065] The input coupling mechanism IN is the same as 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 transmission mechanism FA0603 is connected to the output unit G10. The output unit G10 includes the transmission mechanism output gear G11 and the output gear set G12. The output gear set G12 drives a differential (not shown). The differential (not shown) includes a final reduction gear.

[0067] The rotation of the carrier C46 of PG46 is transmitted to the output gear G11 of the transmission mechanism, and this rotation is then transmitted sequentially to the output gear set G12 and the differential (not shown).

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

[0069] The gear ratio iR1 in reverse 1st 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, since the ring gear rotation speed ωR = 0, from equation (2) above, the carrier rotation speed ωC = (-λ) / (1-λ)·ωS, and the reduction ratio JA = (1-λ) / (-λ) = (1-36 / 90) / (-36 / 90) = -1.500. Similarly, calculating, the reduction ratio JR1 for the double pinion planetary gear PGR1 is -3.292, and the reduction ratio JR2 for the double pinion planetary gear PGR2 is -1.000. Therefore, the gear ratio in reverse 1st gear is: iR1 = JA × JR1 × JR2 = (-1.500) × (-3.292) × (-1.000) ≈ -4.94.

[0070] <2nd reverse gear> As shown in Figure 8B, in reverse gear 2 (R2), brakes BR1, BR2, and BB are engaged. The rotation from the input coupling mechanism IN enters the sun gear SB of the double pinion planetary gear PGB, and since the ring gear RB is stopped, the carrier CB rotates. The carrier CA also rotates together with the carrier CB. The rotation of the carrier CA is input to the sun gear SR1 of the double pinion planetary gear PGR1. The following explanation is the same as in reverse gear 1.

[0071] The gear ratio iR2 in reverse 2nd 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, since the ring gear rotation speed ωR = 0, from equation (2) above, 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 in 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 reverse 3rd gear (R3), brakes BR1, BR2, and BC are engaged. The rotation from the input coupling mechanism IN enters carrier CC of the double pinion planetary gear PGC, and since the ring gear RC is stopped, the sun gear SC rotates. Along with the sun gear SC, the carriers CB and CA of the double pinion planetary gears PGB and PGA rotate, and the rotation of carrier CA is input to the sun gear SR1 of the double pinion planetary gear PGR1. The following explanation is the same as in reverse 1st gear.

[0073] The gear ratio iR3 in reverse 3rd 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, since the ring gear rotation speed ωR = 0, from equation (2) above, 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 in reverse 3rd gear is, iR3 = JC × JR1 × JR2 = (-0.667) × (-3.292) × (-1.000) ≈ -2.20.

[0074] <In 1st gear> As shown in Figure 8B, in first gear, brakes B13 and BA are engaged. The rotation from the input coupling mechanism IN enters the sun gear SA of the double pinion planetary gear PGA, and since the ring gear RA is stopped, the carrier CA rotates. The rotation of carrier CA is input to the sun gear S13 of the double pinion planetary gear PG13. Since the ring gear R13 is stopped, the rotation of the sun gear S13 is transmitted to carrier C13. Carrier C46 of the double pinion planetary gear PG46 rotates together with carrier C13 of the double pinion planetary gear PG13. The rotation of carrier C46 is output to output unit G10. In this way, when moving forward, brakes BR1 and BR2 are not engaged for the reverse double pinion planetary gears PGR1 and PGR2, resulting in a completely free state.

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

[0076] <In 2nd gear> As shown in Figure 8B, in 2nd gear (circled number 2), brakes B13 and BB are engaged. The rotation from the input coupling mechanism IN enters the sun gear SB of the double pinion planetary gear PGB, and since 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 process from there is the same as in 1st gear.

[0077] The gear ratio i2 in 2nd gear is calculated as follows. Here, the driving force is transmitted in the order of double pinion planetary gears PGB and PG13. As mentioned above, the reduction ratio JB = -1.000 for the double pinion planetary gear PGB. As mentioned above, the reduction ratio J13 = -3.286 for the double pinion planetary gear PG13. Therefore, the gear ratio in 2nd gear is, i2 = JB × J13 = (-1.000) × (-3.286) ≈ 3.29.

[0078] <In 3rd gear> As shown in Figure 8B, in 3rd gear (circled number 3), brakes B13 and BC are engaged. The rotation from the input coupling mechanism IN enters the carrier CC of the double pinion planetary gear PGC, and since the ring gear RC is stopped, the sun gear SC rotates. The sun gear SC rotates together with the carriers CB and CA of the double pinion planetary gears PGB and PGA, and the rotation of carrier CA is input to the sun gear S13 of the double pinion planetary gear PG13. The process from there is the same as in 1st gear.

[0079] The gear ratio i3 in 3rd gear is calculated as follows. Here, the driving force is transmitted in the order of double pinion planetary gears PGC and PG13. As mentioned above, the reduction ratio JC for the double pinion planetary gear PGC is -0.667. As mentioned above, the reduction ratio J13 for the double pinion planetary gear PG13 is -3.286. Therefore, the gear ratio in 3rd gear is, i2 = JC × J13 = (-0.667) × (-3.286) ≈ 2.20.

[0080] <In 4th gear> As shown in Figure 8B, in 4th gear (circled number 4), brakes B46 and BA are engaged. The rotation from the input coupling mechanism IN enters the sun gear SA of the double pinion planetary gear PGA, and since 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. Since the ring gear R46 is stopped, the rotation of sun gear S46 is transmitted to carrier C46. The rotation of carrier C46 is output to output unit G10.

[0081] The gear ratio i4 in 4th gear is calculated as follows. Here, the driving force is transmitted in the order of double pinion planetary gears PGA and PG46. For the double pinion planetary gear PGA, the reduction ratio JA = -1.500 is as stated above. For the double pinion planetary gear PG46, since the ring gear rotation speed ωR = 0, from equation (2) above, 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 in 4th gear is, i4 = JA × J46 = (-1.500) × (-1.000) ≈ 1.50.

[0082] <At 5th gear> As shown in Figure 8B, in 5th gear (circled number 5), brakes B46 and BB are engaged. The rotation from the input coupling mechanism IN enters the sun gear SB of the double pinion planetary gear PGB, and since the ring gear RB is stopped, the carrier CB rotates. Along with the carrier CB, the carrier CA of the double pinion planetary gear PGA rotates, and the rotation of carrier CA is input to the sun gear S46 of the double pinion planetary gear PG46. The process from there 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 in the order of double pinion planetary gear PGB, PG46. For the double pinion planetary gear PGB, the reduction ratio JB = -1.000 is as stated above. For the double pinion planetary gear PG46, the reduction ratio J46 = -1.000 is as stated above. Therefore, the gear ratio in 5th gear is, i5 = JB × J46 = (-1.000) × (-1.000) ≈ 1.00.

[0084] <At 6th gear> As shown in Figure 8B, in 6th gear (circled number 6), brakes B46 and BC are engaged. The rotation from the input coupling mechanism IN enters carrier CC of the double pinion planetary gear PGC, and since the ring gear RC is stopped, the sun gear SC rotates. The sun gear SC rotates together with the carriers CB and CA of the double pinion planetary gears PGB and PGA, and the rotation of carrier CA is input to the sun gear S46 of the double pinion planetary gear PG46. The process from there is the same as in 4th 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 gears PGC and PG46. As mentioned above, the reduction ratio JC = -0.667 for the double pinion planetary gear PGC. As mentioned above, the reduction ratio J46 = -1.000 for the double pinion planetary gear PG46. Therefore, the gear ratio in 6th gear is, i6 = JC × J46 = (-0.667) × (-1.000) ≈ 0.67.

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

[0087] Regarding the first planetary gear group G1, as shown in Figure 8B, in each gear position, when one of the brakes BA, BB, or BC is engaged, one of the double pinion planetary gears PGA, PGB, or PGC is activated. Both double pinion planetary gears PGA and PGB are sun gear inputs and carrier outputs, while double pinion planetary gear PGC is a carrier input and sun gear output. Therefore, in any case when one of the double pinion planetary gears PGA, PGB, or PGC is activated, the rotation of the output is reversed relative to the input. Consequently, the first planetary gear group G1 as a group has a reverse output function.

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

[0089] Regarding the third planetary gear group G3, as shown in Figure 8B, in forward gears 1-3, the double pinion planetary gear PG13 is activated by the engagement of brake B13, and in forward gears 4-6, the double pinion planetary gear PG46 is activated by the engagement of brake B46. Both double pinion planetary gears PG13 and PG46 are sun gear inputs and carrier outputs. Therefore, when either double pinion planetary gear PG13 or PG46 is activated, the rotation of the output is reversed relative to the input. Consequently, the third planetary gear group G3 as a group has a reverse output function.

[0090] In the above configuration, when moving forward, the first planetary gear group G1 and the third planetary gear group G3, which are connected in series, are used. As mentioned above, both the first planetary gear group G1 and the third planetary gear group G3 have a reverse output function, so they produce a forward output as a transmission (reverse + reverse = forward). Also, when moving backward, the first planetary gear group G1 and the second planetary gear group G2, which are connected in series, are used. As mentioned above, the first planetary gear group G1 has a reverse output function, and the second planetary gear group G2 has a forward output function, so they produce a reverse output as a transmission (reverse + forward = reverse).

[0091] <Effects> In this embodiment, the operation of each planetary gear PGA, PGB, PGC, PG13, PG46, PGR1, and PGR2 is performed solely by brakes, without the use of clutches. This allows for simplification, miniaturization, weight reduction, and cost reduction of the structure, and improves the overall layout flexibility of the automatic transmission.

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

[0093] Furthermore, in this embodiment, since there are multiple reverse gear stages, it is possible to use a gear stage that provides appropriate driving force even when reversing. Moreover, since only brakes are used and no clutch is used, the overall structure of the AT is compact and has a high degree of layout flexibility, making it easier to realize multiple reverse gear stages from a layout perspective.

[0094] Furthermore, in this embodiment, two planetary gear groups (G1, G3) having a reverse output function are connected in series to function as a transmission with forward rotation output. This provides an output rotation in the same direction as the input rotation direction when moving forward. Additionally, a first planetary gear group G1 having a reverse output function and a second planetary gear group G2 having a forward rotation output function are connected in series to function as a transmission with reverse output. This provides an output rotation in the opposite direction to the input rotation direction when moving backward.

[0095] <Third Embodiment> The RB0803 transmission mechanism for a stepped automatic transmission for automobiles according to a third embodiment of the present invention will be described with reference to Figures 9A and 9B. Figure 9A is a skeleton diagram of the transmission mechanism for an 8-speed forward and 2-speed reverse automatic transmission for rear-wheel drive. For simplicity, only one side (upper half) of the rotating body is shown with the rotation axis X as the center.

[0096] The RB0803 transmission mechanism features five sets of double pinion planetary gears PGZ, PG11, PG12, PG21, PG22 and a Ravigno planetary gear PG0.

[0097] The double pinion planetary gears PGZ, PG11, PG12, PG21, PG22 and the Ravigno planetary gear PG0 each have known configurations. 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 Ravigno planetary gear PG0 consists of two sun gears S1 and S2, a ring gear R0, and a carrier C0.

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

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

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

[0101] <At reverse 1st gear> As shown in Figure 9B, in reverse gear 1 (R1), brakes B21 and B01 are engaged. 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. Carrier C21 of the double pinion planetary gear PG21 rotates together with carriers CZ, C11, and C12, and since brake B21 is engaged, the rotation is transmitted to the sun gear S21. The rotation of sun gear S21 is input to the sun gear S2 of the Ravigno planetary gear PG0. Since carrier C0 is stopped, the rotation of sun gear S2 is transmitted to the ring gear R0. The rotation of ring gear R0 is output to the output section OUT.

[0102] For the double pinion planetary gear PGZ, since the ring gear rotation speed ωR=0, from (Equation 2) above, the carrier rotation speed ωC=-1.000ωS, and the reduction ratio JZ=-1.000. That is, the uppermost double pinion planetary gear PGZ inputs a reverse constant speed output to the downstream gear shifting 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 Ravigno planetary gear PG0, in the above equation (3B), the carrier is fixed (ωC=0), ωR = -β·ωS2 = (-30 / 90)·ωS2 = (-0.333)·ωS2 Therefore, the reduction ratio J0 = -3.000. Consequently, the gear ratio in reverse 1st gear is: iR1 = JZ × J21 × J0 = (-1.000) × (-1.000) × (-3.000) ≈ -3.00.

[0103] <2nd reverse gear> As shown in Figure 9B, in reverse 2nd gear (R2), brakes B22 and B01 are engaged. 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. Carrier C22 of the double pinion planetary gear PG22 rotates together with carriers CZ, C11, C12, and C21, and since 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 Ravigno planetary gear PG0. Since 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 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 (Equation 2) above, (1 - 39 / 90)·ωC = (-39 / 90)·ωS, so the reduction ratio J22 = -0.765. For the Ravigno planetary gear PG0, the reduction ratio J0 = -3.000, as above. Therefore, the gear ratio in reverse 2nd gear is: iR2 = JZ × J22 × J0 = (-1.000) × (-0.765) × (-3.000) ≈ -2.29.

[0105] <When in forward 1st gear> As shown in Figure 9B, in forward 1st gear (circled number 1), 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 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 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 Ravigno planetary gear PG0. Since 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 (Equation 2) above, (1 - 60 / 105)·ωC = (-60 / 105)·ωS, so the reduction ratio J11 = -1.333. For the Ravigno planetary gear PG0, in Equation (3A) above, the carrier is fixed (ωC = 0), so ωR = α·ωS1 = (24 / 90)·ωS1 = (0.267)·ωS1, and the reduction ratio J0 = 3.750. Therefore, the gear ratio in first forward gear is: i1 = JZ × J11 × J0 = (-1.000) × (-1.333) × (3.750) ≈ 5.00.

[0107] <When moving forward in 2nd gear> As shown in Figure 9B, in forward 2nd gear (circled number 2), brakes B12 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 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 carriers CZ and C11, and since 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 Ravigno planetary gear PG0. Since 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 (Equation 2) above, (1 - 42 / 90)·ωC = (-42 / 90)·ωS, so the reduction ratio J12 = -0.875. For the Ravigno planetary gear PG0, the reduction ratio J0 = 3.750, as above. Therefore, the gear ratio in forward 2nd gear is: i2 = JZ × J12 × J0 = (-1.000) × (-0.875) × (3.750) ≈ 3.28.

[0109] <When in 3rd gear forward> As shown in Figure 9B, in forward 3rd gear (circled number 3), 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 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 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 Ravigno planetary gear PG0. Since 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 is the same as above. For the double pinion planetary gear PG11, the reduction ratio J11 = -1.333 is the same as above. For the Ravigno planetary gear PG0, since sun gear S1 is the input, sun gear S2 is fixed, and ring gear R0 is the output, if we set 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. Alternatively, Figure 4(b) may be used directly. Therefore, the gear ratio in forward 3rd gear is i3 = JZ × J11 × J0 = (-1.000) × (-1.333) × (1.688) ≈ 2.25.

[0111] <When in 4th gear forward> As shown in Figure 9B, in forward 4th gear (circled number 4), 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 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 carriers CZ and C11, and since 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 Ravigno planetary gear PG0. Since 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 Ravigno planetary gear PG0, the reduction ratio J0 = 1.688, as above. Therefore, the gear ratio in forward 4th gear is: i4 = JZ × J12 × J0 = (-1.000) × (-0.875) × (1.688) ≈ 1.48.

[0113] <When moving forward in 5th gear> As shown in Figure 9B, in forward 5th gear (circled number 5), brakes B11 and B21 are engaged. 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. Carrier C11 of the double pinion planetary gear PG11 rotates together with carrier CZ, and since brake B11 is engaged, the rotation is transmitted to sun gear S11. The rotation of sun gear S11 is input to sun gear S1 of the Ravigno planetary gear PG0. Furthermore, carrier C21 of the double pinion planetary gear PG21 rotates together with carrier CZ, carrier C11, and carrier C12, and since brake B21 is engaged, the rotation is transmitted to sun gear S21. The rotation of sun gear S21 is input to sun gear S2 of the Ravigno planetary gear PG0. In other words, the Ravigno planetary gear PG0 has two input systems (sun gears S1 and S2). The rotation of sun gears S1 and S2 is transmitted to the ring gear R0 via carrier C0. The rotation of 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 above. For the double pinion planetary gear PG11, the reduction ratio J11 = -1.333, as above. For the double pinion planetary gear PG21, the reduction ratio J21 = -1.000, as above. For the Ravigno planetary gear PG0, by eliminating ωC from equations (3A) and (3B) above and solving for ωR, we get 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}]. Here, J1 is the reduction ratio of the input to sun gear S1, and J2 is the reduction ratio of the input to sun gear S2. Alternatively, Figure 4(c) may be used. Therefore, the gear ratio in 5th forward gear is, i5= JZ×J0 = (-1.000)×(α+β) / [{α(1+β) / (-1.333)}+{β(1-α) / (-1.000)}]≒1.17.

[0115] <When moving forward in 6th gear> As shown in Figure 9B, in forward 6th gear (circled number 6), 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 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 Ravigno 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 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 Ravigno planetary gear PG0. The rotation of sun gears S1 and S2 is transmitted to ring gear R0 via carrier C0. The rotation of ring gear R0 is output to output unit OUT.

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

[0117] <At 7th gear forward> As shown in Figure 9B, in forward 7th gear (circled number 7), brakes B12 and B21 are engaged. 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. Carrier C12 of the double pinion planetary gear PG12 rotates together with carriers CZ and C11, and since brake B12 is engaged, the rotation is transmitted to sun gear S12. The rotation of sun gear S12 is input to sun gear S1 of the Ravigno planetary gear PG0. Furthermore, carrier C21 of the double pinion planetary gear PG21 rotates together with carriers CZ, C11 and C12, and since brake B21 is engaged, the rotation is transmitted to sun gear S21. The rotation of sun gear S21 is input to sun gear S2 of the Ravigno planetary gear PG0. The rotation of sun gears S1 and S2 is transmitted to ring gear R0 via carrier C0. The rotation of ring gear R0 is output to output unit OUT.

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

[0119] <When moving forward in 8th gear> As shown in Figure 9B, in forward 8th gear (circled number 8), brakes B12 and B22 are engaged. 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. Carrier C12 of the double pinion planetary gear PG12 rotates together with carriers CZ and C11, and since brake B12 is engaged, the rotation is transmitted to sun gear S12. The rotation of sun gear S12 is input to sun gear S1 of the Ravigno planetary gear PG0. Furthermore, carrier C22 of the double pinion planetary gear PG22 rotates together with carriers CZ, C11, C12 and C21, and since brake B22 is engaged, the rotation is transmitted to sun gear S22. The rotation of sun gear S22 is input to sun gear S2 of the Ravigno planetary gear PG0. The rotation of sun gears S1 and S2 is transmitted to ring gear R0 via carrier C0. The rotation of ring gear R0 is output to output unit OUT.

[0120] For the double pinion planetary gear PGZ, the reduction ratio JZ = -1.000 is obtained, as above. For the double pinion planetary gear PG12, the reduction ratio J12 = -0.875 is obtained, as above. For the double pinion planetary gear PG22, the reduction ratio J22 = -0.765 is obtained, as above. For the Ravigno planetary gear PG0, the reduction ratio J0 = (α + β) / [{α(1 + β) / J1} + {β(1 - α) / J2}] is obtained, as above. Therefore, the gear ratio in 8 forward gears 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 (S1 x 2) to the sun gear S1 of the Ravigno planetary gear PG0. Specifically, 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 brakes B11 and B12. Also, there are two inputs (S2 x 2) to the sun gear S2 of the Ravigno planetary gear PG0. Specifically, 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 brakes B21 and B22. There is no input to the carrier C0 of the Ravigno planetary gear PG0.

[0122] <Effects> In this embodiment, the operation of each planetary gear PGZ, PG11, PG12, PG21, PG22, and PG0 is performed solely by brakes, without the use of clutches. This simplifies the structure, makes it smaller and lighter, reduces costs, and improves the overall layout flexibility of the automatic transmission. Furthermore, because it has multiple reverse gear stages, it is possible to use a gear stage with appropriate driving force even when reversing. Moreover, because only brakes are used without clutches, the overall structure of the automatic transmission is compact and offers great layout flexibility, making it easier to implement multiple reverse gear stages from a layout perspective.

[0123] Furthermore, in this embodiment, two shift speeds are input to the sun gear S1 of the Ravigno planetary gear PG0, and two shift speeds are also input to the sun gear S2. Specifically, 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 Ravigno 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 brakes B11 and B12. In addition, 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 Ravigno planetary gear PG0. Furthermore, the input from the double pinion planetary gear PG21 and the input from the double pinion planetary gear PG22 are switched so that either one is input to the sun gear S2 by the operation of brakes B21 and B22. In other words, multiple gear speeds are input to each of the two input members (sun gears S1 and S2) of the composite planetary gear, allowing for an increase in gear ratios with a simple structure.

[0124] <Fourth Embodiment> The FA0501 transmission mechanism for a stepped automatic transmission for automobiles according to the fourth embodiment of the present invention will be described with reference to Figures 10A and 10B. Figure 10A is a skeleton diagram of the transmission mechanism for a 5-speed forward and 5-speed reverse automatic transmission for front-wheel drive. To avoid complexity, only one side (upper half) of the rotating body is shown with the rotation axis X as the center.

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

[0126] The FA0501 transmission mechanism is further equipped with brakes BF, B1, B2, B3, B4, and B5. Each of these brakes is a wet multi-plate type, with rotating discs (the side being braked) and stationary discs packed alternately, and the stationary discs are 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 applies to brakes B2, B3, B4, and B5.

[0127] <In 1st gear> As shown in Figure 10B, in 1st gear (circled number 1), brakes BF and B1 are engaged. The rotation from the input coupling mechanism IN enters the sun gear SF of the double pinion planetary gear PGF, and since 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. Since 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 unit G0.

[0128] For the double pinion planetary gear PGF, the ring gear rotation speed ωR = 0, so from equation (2) above, 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 equation (2) above, 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 Figure 10B, in reverse gear 1 (R1), the clutch hub sleeve CHS and brake B1 are engaged. That is, the sleeve SL of the clutch hub sleeve CHS slides (shown by the dotted line in Figure 10A) and engages with the carrier CF of the double pinion planetary gear PGF. At this time, the carrier CF rotates at the same speed as the input IN. In other words, it is directly connected. The rotation of the carrier CF is then 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 unit G0.

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

[0131] Thus, the double pinion planetary gear PGF operates during forward movement. Since the tooth ratio λF = 45 / 90 of the double pinion planetary gear PGF, the reduction ratio is -1.000, and the input to the variable speed planetary gear group (double pinion planetary gears PG1 to PG5) is constant speed reverse rotation. On the other hand, during reverse movement, the carrier CF of the double pinion planetary gear PGF is directly connected to the input IN by the engagement of the clutch hub sleeve CHS, so the input to the variable speed planetary gear group (double pinion planetary gears PG1 to PG5) is constant speed forward rotation. In this embodiment, the clutch hub sleeve CHS is used as the reverse gear selection method.

[0132] For gears 2 through 5, the only difference is that double pinion planetary gears PG2 through PG5 operate instead of double pinion planetary gear PG1, which operates in gear 1, so the explanation is omitted. Similarly, for reverse gears 2 through 5, the only difference is that double pinion planetary gears PG2 through PG5 operate instead of double pinion planetary gear PG1, which operates in reverse gear 1, so the explanation is omitted.

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

[0134] In this embodiment, a clutch hub sleeve CHS is used as the reverse gear selection method, allowing the reverse gear to be implemented with a simple mechanism.

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

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

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

[0138] Figure 12 shows a skeleton diagram and brake application table of the FA0602 transmission mechanism according to another embodiment. The FA0602 transmission mechanism is a 6-speed forward and 3-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single shaft. It has a DPPG 7-set configuration. The reverse power flow consists of 2 DPPG sets, with input and output rotating in the same direction. When moving forward, the reverse DPPG is completely free.

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

[0140] Figure 14 shows a skeleton diagram and brake application table for the FA0605 transmission mechanism according to another embodiment. The FA0605 transmission mechanism is a 6-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single shaft. It has a DPPG 7-set configuration. The reverse power flow consists of 1 DPPG set, with input and output rotating in the same direction. The 1st and 2nd reverse gears and the 1st and 2nd forward gears use an additional reduction system. The DPPG for additional reduction rotates in the same direction for input and output.

[0141] The additional reduction method refers to the process of applying additional reduction to the low-speed planetary gears because the desired gear ratio for the automatic transmission cannot be achieved using only the gear ratios of the low-speed shifting planetary gears. In the FA0605 transmission mechanism, this corresponds to Brake BZ in the Brake Application Table. The reduction ratio for DPPG with Brake BR activated is 1.765, and the reduction ratio for DPPG with Brake B12 activated is -1.727. Since this is insufficient to achieve the desired reduction ratio for the automatic transmission, the reduction ratio of 2.000 for Brake BZ activation is additionally applied.

[0142] Figure 15 shows a skeleton diagram and brake application table for the FA0606 transmission mechanism according to another embodiment. The FA0606 transmission mechanism is a 6-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single shaft. It is configured as a combination of DPPG6 sets and CHS. CHS is applied to the reverse power flow. The 1st and 2nd reverse and 1st and 2nd forward gears use an additional reduction method. The DPPG for additional reduction rotates in the same direction for input and output.

[0143] Figure 16 shows a skeleton diagram and brake application table of the FA0801 transmission mechanism according to another embodiment. The FA0801 transmission mechanism is an 8-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single shaft. It consists of a combination of 5 sets of double pinion planetary gears (DPPG) and 2 sets of single pinion planetary gears (SPPG).

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

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

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

[0147] Figure 20 shows a skeleton diagram and brake application table for the FA0805 transmission mechanism according to another embodiment. The FA0805 transmission mechanism is an 8-speed forward and 4-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It has a DPPG 9 set configuration. Reverse 1-4 and forward 1-4 use an additional reduction method. The DPPG for additional reduction rotates in the same direction for input and output.

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

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

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

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

[0152] Figure 25 shows a skeleton diagram and brake application table of the FB0601 transmission mechanism according to another embodiment. The FB0601 transmission mechanism is a 6-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It is configured as a combination of DPPG4 sets and Ravigno-type planetary gears (Ravigno PG).

[0153] Figure 26 shows a skeleton diagram and brake application table of the FB0602 transmission mechanism according to another embodiment. The FB0602 transmission mechanism is a 6-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It is configured as a combination of DPPG3 set and Ravigno PG.

[0154] Figure 27 shows a skeleton diagram and brake application table of the FB0603 transmission mechanism according to another embodiment. The FB0603 transmission mechanism is a 6-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It is configured as a combination of the DPPG3 set and the Ravigno PG.

[0155] Figure 28 shows a skeleton diagram and brake application table of the FB0604 transmission mechanism according to another embodiment. The FB0604 transmission mechanism is a 6-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It is configured as a combination of DPPG3 sets and Ravigno PG. The DPPG for the S1 and S2 inputs of the Ravigno PG is an enlarged version.

[0156] Figure 29 shows a skeleton diagram and brake application table of the FB0801 transmission mechanism according to another embodiment. The FB0801 transmission mechanism is an 8-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It is configured as a combination of DPPG4 sets and Ravigno PG. The inputs to the Ravigno PG are S1 x 2 (two inputs to sun gear S1), S2 x 2 (two inputs to sun gear S2), and there is no input to C (carrier). The same applies below.

[0157] Figure 30 shows a skeleton diagram and brake application table of the FB0802 transmission mechanism according to another embodiment. The FB0802 transmission mechanism is an 8-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It is configured as a combination of DPPG4 sets and Ravigno PG. The inputs to the Ravigno PG are S1 x 3, S2 x 1, and no input for C.

[0158] Figure 31 shows a skeleton diagram and brake application table of the FB0901 transmission mechanism according to another embodiment. The FB0901 transmission mechanism is a 9-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It is configured as a combination of DPPG4 sets and Ravigno PG. The inputs to the Ravigno PG are S1 x 3, S2 x 1, and no input for C.

[0159] Figure 32 shows a skeleton diagram and brake application table of the FB1001 transmission mechanism according to another embodiment. The FB1001 transmission mechanism is a 10-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on a single axle. It is configured as a combination of 4 DPPG sets and a Ravigno PG. The inputs to the Ravigno PG are S1 x 2, S2 x 1, and C x 1. One of the two DPPGs for the S1 input of the Ravigno PG is an enlarged version.

[0160] Figure 33 shows a skeleton diagram and brake application table for the FC0601 transmission mechanism according to another embodiment. The FC0601 transmission mechanism is a 6-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged in a 3-axis configuration (#1, #2, #3). It is configured as a combination of DPPG6 sets and CHS. It has 2-axis output, and there are two types of final gear pinions: high-speed (FGP-Hi) and low-speed (FGP-Lo). CHS is applied to the reverse power flow.

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

[0162] Figure 35 shows a skeleton diagram and brake application table of the FC0603 transmission mechanism according to another embodiment. The FC0603 transmission mechanism is a 6-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on two axes (#1, #2). It is configured as a combination of DPPG4 sets and SPPG2 sets. Two types of transfer gear trains TG are available. For rotational transmission between axle #1 and axle #2, TG1 is reverse transmission and TG2 is forward rotation transmission.

[0163] Figure 36 shows a skeleton diagram and brake application table for the FC0604 transmission mechanism according to another embodiment. The FC0604 transmission mechanism is a 6-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on two axes (#1, #2). It has a DPPG 7-set configuration. The reverse power flow has a DPPG 2-set configuration. When moving forward, the reverse DPPG is completely free.

[0164] Figure 37 shows a skeleton diagram and brake application table for the FC0801 transmission mechanism according to another embodiment. The FC0801 transmission mechanism is an 8-speed forward and 4-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged in a 3-axis configuration (#1, #2, #3). It has an 8-set DPPG configuration. It has 2-axis output. The reverse power flow has a 2-set DPPG configuration. When moving forward, the reverse DPPG is completely free.

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

[0166] Figure 39 shows a skeleton diagram and brake application table for the FC0901 transmission mechanism according to another embodiment. The FC0901 transmission mechanism is a 9-speed forward and 3-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged in a 3-axis configuration (#1, #2, #3). It has an 8-set DPPG configuration. It has 2-axis output, and there are two types of final gear pinions: one for high speed and one for low speed. The reverse power flow has a 2-set DPPG configuration. When moving forward, the reverse DPPG is completely free.

[0167] Figure 40 shows a skeleton diagram and brake application table of the FC0902 transmission mechanism according to another embodiment. The FC0902 transmission mechanism is a 9-speed forward and 3-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on two axes (#1, #2). It has an 8-set DPPG configuration. The reverse power flow has a 2-set DPPG configuration. When moving forward, the reverse DPPG is completely free.

[0168] Figure 41 shows a skeleton diagram and brake application table of the FD0601 transmission mechanism according to another embodiment. The FD0601 transmission mechanism is a 6-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on two axes. It is a combination of DPPG3 sets and Ravigno PG. The transfer gear train TG has two types (TG1, TG2) for reduction and constant speed, and is arranged adjacent to each other.

[0169] Figure 42 shows a skeleton diagram and brake application table of the FD0801 transmission mechanism according to another embodiment. The FD0801 transmission mechanism is an 8-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on two axes. It is configured as a combination of DPPG4 sets and Ravigno PG. The transfer gear train TG is arranged adjacently at constant velocity. The inputs to the Ravigno PG are S1×2, S2×2, and there is no input for C.

[0170] Figure 43 shows a skeleton diagram and brake application table of the FD0901 transmission mechanism according to another embodiment. The FD0901 transmission mechanism is a 9-speed forward and 3-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on two axes. It is configured as a combination of DPPG6 set and Ravigno PG.

[0171] Figure 44 shows a skeleton diagram and brake application table of the FD0902 transmission mechanism according to another embodiment. The FD0902 transmission mechanism is a 9-speed forward and 2-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on two axes. It is configured as a combination of DPPG5 sets and Ravigno PG. The inputs to the Ravigno PG are S1×2, S2×2, and C×1.

[0172] Figure 45 shows a skeleton diagram and brake application table of the FD0903 transmission mechanism according to another embodiment. The FD0903 transmission mechanism is a 9-speed forward and 1-speed reverse transmission mechanism for front-wheel drive. The main gear train is arranged on two axes. It is configured as a combination of DPPG4 sets and Ravigno PG. The inputs to the Ravigno PG are S1×2, S2×1, and C×1.

[0173] Figure 46 shows a skeleton diagram and brake application table of the RA0801 transmission mechanism according to another embodiment. The RA0801 transmission mechanism is an 8-speed forward and 4-speed reverse transmission mechanism for rear-wheel drive. It has an 8-set DPPG configuration. The reverse power flow consists of 2 DPPG sets, with input and output rotating in the same direction. When moving forward, the reverse DPPG is completely free.

[0174] Figure 47 shows a skeleton diagram and brake application table of the RA0802 transmission mechanism according to another embodiment. The RA0802 transmission mechanism is an 8-speed forward and 4-speed reverse transmission mechanism for rear-wheel drive. It consists of a DPPG6 set and an SPPG1 set.

[0175] Figure 48 shows a skeleton diagram and brake application table of the RA0803 transmission mechanism according to another embodiment. The RA0803 transmission mechanism is an 8-speed forward and 4-speed reverse transmission mechanism for rear-wheel drive. It is configured as a combination of DPPG6 sets, SPPG1 set, and CHS. The SPPG has a fixed reduction function for reverse and outputs reverse power flow when the CHS is activated.

[0176] Figure 49 shows a skeleton diagram and brake application table of the RA0804 transmission mechanism according to another embodiment. The RA0804 transmission mechanism is an 8-speed forward and 4-speed reverse transmission mechanism for rear-wheel drive. It is configured as a combination of DPPG6 sets, SPPG1 set, and CHS. The SPPG provides deceleration output for forward gears 1-4 and reverse gears 1-4. When driving in forward gears 5-8, the SPPG is completely free.

[0177] Figure 50 shows a skeleton diagram and brake application table of the RA0805 transmission mechanism according to another embodiment. The RA0805 transmission mechanism is an 8-speed forward and 2-speed reverse transmission mechanism for rear-wheel drive. It is configured as a combination of a DPPG7 set and a CHS. The uppermost DPPG has a forward rotation fixed reduction function dedicated to reverse, and outputs reverse power flow when the CHS is activated.

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

[0179] Figure 52 shows a skeleton diagram and brake application table of the RA0902 transmission mechanism according to another embodiment. The RA0902 transmission mechanism is a rear-wheel drive transmission mechanism with 9 forward speeds and 3 reverse speeds. It consists of DPPG6 sets and SPPG1 sets.

[0180] Figure 53 shows a skeleton diagram and brake application table of the RA0903 transmission mechanism according to another embodiment. The RA0903 transmission mechanism is a rear-wheel drive transmission mechanism with 9 forward speeds and 3 reverse speeds. It consists of DPPG6 sets and SPPG1 sets.

[0181] Figure 54 shows a skeleton diagram and brake application table of the RA1001 transmission mechanism according to another embodiment. The RA1001 transmission mechanism is a 10-speed forward and 5-speed reverse transmission mechanism for rear-wheel drive. It is configured as a combination of DPPG7 set, SPPG1 set, and CHS. The SPPG has a fixed reduction function for reverse and outputs reverse power flow when the CHS is activated.

[0182] Figure 55 shows a skeleton diagram and brake application table of the RA1201 transmission mechanism according to another embodiment. The RA1201 transmission mechanism is a 12-speed forward and 3-speed reverse transmission mechanism for rear-wheel drive. It has a DPPG9 set configuration. The reverse power flow consists of DPPG2 sets, with input and output rotating in the same direction. When moving forward, the reverse DPPG is in a completely free state.

[0183] Figure 56 shows a skeleton diagram and brake application table of the RB0601 transmission mechanism according to another embodiment. The RB0601 transmission mechanism is a 6-speed forward and 1-speed reverse transmission mechanism for rear-wheel drive. It is configured as a combination of DPPG4 sets and Ravigno PG. The uppermost DPPG inputs the reverse deceleration output to the transmission function unit.

[0184] Figure 57 shows a skeleton diagram and brake application table of the RB0801 transmission mechanism according to another embodiment. The RB0801 transmission mechanism is an 8-speed forward and 1-speed reverse transmission mechanism for rear-wheel drive. It is configured as a combination of DPPG5 sets and Ravigno PG. The uppermost DPPG inputs the reverse deceleration output to the transmission function unit. The inputs to the Ravigno PG are S1×2, S2×1, and C×1.

[0185] Figure 58 shows a skeleton diagram and brake application table of the RB0802 transmission mechanism according to another embodiment. The RB0802 transmission mechanism is an 8-speed forward and 1-speed reverse transmission mechanism for rear-wheel drive. It is configured as a combination of DPPG5 sets and Ravigno PG. The uppermost DPPG inputs the reverse deceleration output to the transmission function unit. The inputs to the Ravigno PG are S1 x 3, S2 x 1, and no input for C.

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

[0187] Figure 60 shows a skeleton diagram and brake application table of the RB0902 transmission mechanism according to another embodiment. The RB0902 transmission mechanism is a rear-wheel drive transmission mechanism with 9 forward and 2 reverse speeds. It is configured as a combination of DPPG6 sets and a Ravigno type PG. The uppermost DPPG inputs the reverse deceleration output to the transmission function unit. The inputs to the Ravigno type PG are S1×2, S2×2, and C×1.

[0188] Figure 61 shows a skeleton diagram and brake application table of the RB0903 transmission mechanism according to another embodiment. The RB0903 transmission mechanism is a rear-wheel drive transmission mechanism with 9 forward and 2 reverse speeds. It is configured as a combination of DPPG5 sets and a Ravigno type PG. The two upstream DPPGs input the reverse deceleration output to the transmission function unit. The inputs to the Ravigno type PG are S1×2, S2×2, and C×2.

[0189] Figure 62 shows a skeleton diagram and brake application table for the RB0904 transmission mechanism according to another embodiment. The RB0904 transmission mechanism is a rear-wheel drive transmission mechanism with 9 forward and 3 reverse speeds. It is configured as a combination of DPPG6 sets and a Ravigno type PG. The three upstream DPPGs input reverse deceleration and constant speed output to the transmission function unit. The inputs to the Ravigno type PG are S1×3, S2×3, and C×3.

[0190] Figure 63 shows a skeleton diagram and brake application table of the RB0905 transmission mechanism according to another embodiment. The RB0905 transmission mechanism is a rear-wheel drive transmission mechanism with 9 forward and 3 reverse speeds. It is configured as a combination of a DPPG6 set and a Ravigno type PG. The inputs to the Ravigno type PG are S1 x 4, S2 x 4, and no input for C.

[0191] Figure 64 shows a skeleton diagram and brake application table of the RB1001 transmission mechanism according to another embodiment. The RB1001 transmission mechanism is a 10-speed forward and 1-speed reverse transmission mechanism for rear-wheel drive. It is configured as a combination of a DPPG5 set and a Ravigno type PG. The inputs to the Ravigno type PG are S1×2, S2×1, and C×1.

[0192] FIG. 65 shows a skeleton diagram and a brake application table of a transmission mechanism RB1002 according to another embodiment. The transmission mechanism RB1002 is a forward 10-speed and reverse 1-speed transmission mechanism for rear-wheel drive. It is a combined configuration of a DPPG5 set and a Ravigneaux type PG. The inputs to the Ravigneaux type PG are S1×2, S2×1, and C×1. The last DPPG reversely decelerates the ring gear output (all shift positions) of the Ravigneaux PG to obtain the output of the transmission.

[0193] FIG. 66 shows a skeleton diagram and a brake application table of a transmission mechanism RB1003 according to another embodiment. The transmission mechanism RB1003 is a forward 10-speed and reverse 2-speed transmission mechanism for rear-wheel drive. It is a combined configuration of a DPPG5 set and a Ravigneaux type PG. The upstream two DPPGs input the reverse deceleration and speed increase outputs to the transmission function unit. The inputs from the transmission function unit to the Ravigneaux type PG are S1×2, S2×1, and C has no input.

[0194] FIG. 67 shows a skeleton diagram and a brake application table of a transmission mechanism RB1201 according to another embodiment. The transmission mechanism RB1201 is a forward 12-speed and reverse 2-speed transmission mechanism for rear-wheel drive. It is a combined configuration of a DPPG5 set and a Ravigneaux type PG. The inputs to the Ravigneaux type PG in the transmission function unit are S1×1, S2×1, and C×1. The upstream two DPPGs input the reverse deceleration output to the transmission function unit.

[0195] FIG. 68 shows a skeleton diagram and a brake application table of a transmission mechanism RB1202 according to another embodiment. The transmission mechanism RB1202 is a forward 12-speed and reverse 2-speed transmission mechanism for rear-wheel drive. It is a combined configuration of a DPPG6 set and a Ravigneaux type PG. The inputs from the four DPPGs in the transmission function unit to the Ravigneaux type PG are S1×3, S2×1, and C has no input. The upstream two DPPGs input the reverse deceleration and speed increase outputs to the transmission function unit.

[0196] The transmission mechanism of the stepped automatic transmission for automobiles according to the present invention is not limited to the above-described embodiments, and various modifications and improvements are possible within the scope described in the claims.

Explanation of Signs

[0197] PG11, PG12, PG21, PG22, PGR1, PGR2, PG13, PG46, PGA, PGB, PGC, PGZ, PGF, PG1, PG2, PG3, PG4, PG5... Double pinion planetary gear, PG0... Ravigno 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. Multiple planetary gears are connected to form a gear train. Each of the aforementioned multiple planetary gears comprises a sun gear, a ring gear, and a carrier that rotatably supports multiple pinion gears. A gear shifting mechanism for a stepped automatic transmission for automobiles, characterized in that each of the aforementioned multiple planetary gears is operated solely by brakes without using a clutch.

2. The aforementioned plurality of planetary gears include composite planetary gears, The gear shift mechanism for a stepped automatic transmission for an automobile according to claim 1, characterized in that multiple gear shift speeds are input to one input member of the composite planetary gear.

3. A gear shift mechanism for a stepped automatic transmission for an automobile according to claim 1, characterized by having multiple reverse gear stages.

4. The aforementioned plurality of planetary gears include two single-row double-pinion planetary gears connected in series, The power flow of the reverse gear is configured by the two single-row double-pinion planetary gears. A gear shifting mechanism for a stepped automatic transmission for an automobile according to any one of 1 to 3, characterized in that the input rotation direction and the output rotation direction are the same when the two single-row double-pinion planetary gears are viewed as a group.

5. Equipped with a clutch hub sleeve, A gear shift mechanism for a stepped automatic transmission for an automobile according to any one of claims 1 to 3, characterized in that the clutch hub sleeve is used as the reverse gear selector.

6. The aforementioned plurality of planetary gears are a group of planetary gears consisting of a plurality of single-row planetary gears, and include two groups of planetary gears having a reverse output function. A gear shift mechanism for a stepped automatic transmission for an automobile according to any one of claims 1 to 3, characterized in that two of the aforementioned planetary gear groups are connected in series to provide forward rotation output as a transmission.

7. The plurality of planetary gears includes a first planetary gear group consisting of a plurality of single-row planetary gears and having a reverse output function, and a second planetary gear group consisting of another plurality of single-row planetary gears and having a forward output function. A gear shift mechanism for a stepped automatic transmission for an automobile according to any one of claims 1 to 3, characterized in that the first planetary gear group and the second planetary gear group are connected in series to provide a reverse output as a transmission.

8. The aforementioned plurality of planetary gears include a plurality of single-row planetary gears, A gear shifting mechanism for a stepped automatic transmission for an automobile according to any one of claims 1 to 3, characterized in that, for all of the aforementioned multiple single-row planetary gears, the brake that controls each single-row planetary gear brakes the ring gear of each single-row planetary gear.

Images