JP2008215392A - Automatic transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an automatic transmission which can give a satisfactory acceleration feeling with a clear speed changing feeling in changing a speed by appropriately distributing speed ratios between respective speed changing steps. <P>SOLUTION: The first sun gear S1 and the second carrier C0 of double pinion type planetary gear mechanisms 21, 22 are connected with each other, and are connected to an input shaft 14 so as to transmit power. Also, a first carrier C1 and a first ring gear R1 are connected to a third control brake B-3 and a first control brake B-1 respectively. Further, the third sun gear S2 and the fourth sun gear S3 of single pinion type planetary gear mechanisms 23, 24 are connected with each other, and are disengage-ably connected to the input shaft 14 by a first control clutch C-1. Further, a third ring gear R2 and a fourth ring gear R3 are connected to a second control brake B-2 and a fourth control brake B-4 respectively, and a third carrier C2 is disengage-ably connected to the input shaft 14 by a second control clutch C-2, and also a fourth carrier C3 is connected to an output shaft 17. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to an automatic transmission that shifts rotation of an input shaft to a plurality of stages by a planetary gear device and transmits the speed to an output shaft.

Conventionally, as this type of automatic transmission, for example, an automatic transmission described in Patent Document 1 (hereinafter referred to as “conventional automatic transmission”) is known. That is, in Patent Document 1, a common sun gear directly connected to an input shaft is meshed with a first ring gear via a stepped pinion small-diameter pinion supported by a carrier, and the second ring gear and the stepped pinion large diameter are engaged. The reduction double planetary gear meshed via the pinion, the sun gear of the first single pinion planetary gear and the sun gear of the second single pinion planetary gear are directly connected, and the carrier of the first single pinion planetary gear and the ring gear of the second single pinion planetary gear are directly connected. A first clutch for selectively connecting the transmission-type planetary gear for shifting, a sun gear directly connected to the input shaft and the shifting-type planetary gear for shifting, and a carrier and ring gear directly connected to the input shaft, and the shifting-type planetary gear for shifting. Second clutch to be connected and for deceleration First brake for selectively fixing a first ring gear of a planetary gear, a second brake for selectively fixing a second ring gear of a double planetary gear for reduction, a carrier of a double planetary gear for reduction connected directly, and a first single pinion A third brake for selectively fixing the ring gear of the planetary gear, a fourth brake for selectively fixing the carrier and the ring gear for selective transmission, and an output shaft directly connected to the carrier of the second single pinion planetary gear. An automatic transmission is disclosed in which the rotation of the input shaft is shifted to the eighth forward speed and the reverse speed and transmitted to the output shaft.
JP 2002-213545 A

  By the way, in such an automatic transmission, the increase ratio of the gear ratio (input shaft rotation speed / output shaft rotation speed) when the shift speed is increased by one is called a step ratio. From the viewpoint of obtaining the above, it is preferable that the gears are distributed in a state where there is no great variation among the shift speeds. In addition, if the value of the step ratio itself at each gear stage is too small (that is, a value close to “1”), for example, the number of revolutions within the effective rotation range of the engine at the time of shifting with acceleration. As a result, the feeling of shifting is diminished for the driver, and sufficient acceleration feeling at the time of shifting cannot be obtained.

  In this regard, in the case of the conventional automatic transmission, the step ratio between the fourth shift stage and the fifth shift stage and the step ratio between the fifth shift stage and the sixth shift stage are determined by these shift stages. And the step ratio between the adjacent shift stages on the low speed side and the high speed side are greatly varied. Furthermore, the step ratio between the sixth gear and the seventh gear, which is a high speed gear, is a small step ratio of less than 1.1 that can hardly be expected to bring about a shift feeling. Therefore, in comparison with such a conventional automatic transmission, a gear ratio of eight forward speeds with an appropriately distributed step ratio can be obtained with a clear shift feeling at the time of a shift with acceleration. There has been a need for an automatic transmission with.

  The present invention has been made in view of such circumstances, and an object of the present invention is to provide a clear shift feeling at the time of shifting with acceleration by appropriately allocating the step ratio between the respective gears. It is an object of the present invention to provide an automatic transmission that can obtain a comfortable acceleration feeling.

  In order to achieve the above object, the invention according to claim 1 relating to an automatic transmission is characterized in that a double planetary gear unit for reduction having both a first and second planetary gear mechanisms of a double pinion type and a single pinion both. A double planetary gear device for transmission having a third planetary gear mechanism and a fourth planetary gear mechanism of the type, and in the double planetary gear device for reduction, the first planetary gear mechanism is connected to the first sun gear and the first sun gear. The first planetary gear mechanism includes a first carrier that supports a first pinion that meshes with the first intermediate pinion that meshes with the first pinion, a first ring gear that meshes with the first intermediate pinion, and the second planetary gear mechanism. A second sun gear coupled to the first ring gear; a second carrier that supports a second pinion meshing with the second sun gear and a second intermediate pinion meshing with the second pinion; A second ring gear meshing with the second intermediate pinion is provided, and the first sun gear and the second carrier are connected so as to be able to transmit power to the input shaft, and the first carrier is controlled in a third manner. The first ring gear and the second sun gear are connected to the brake and connected to the first control brake. In the double planetary gear device for speed change, the third planetary gear mechanism includes the third sun gear and the third sun gear. A fourth carrier that supports a third pinion that meshes with the gear; a third ring gear that meshes with the third pinion and is coupled to the second ring gear so as to be capable of transmitting power; and the fourth planetary gear mechanism. A fourth sun gear, a fourth carrier supporting a fourth pinion meshing with the fourth sun gear, and a fourth ring gear meshing with the fourth pinion and connected to the third carrier. The third sun gear and the fourth sun gear are connected to the input shaft so as to be disengageable by a first control clutch, and the third ring gear and the fourth ring gear are connected to the second control brake and the fourth control gear. The third carrier is connected to the control brake, the third carrier is detachably connected to the input shaft by a second control clutch, the fourth carrier is connected to the output shaft, and the rotation of the second ring gear causes the third ring gear to rotate. It is transmitted to the ring gear.

  According to the first aspect of the present invention, the step ratio, which is the increase ratio of the gear ratio (the rotational speed of the input shaft / the rotational speed of the output shaft) when the shift speed is increased by one, is large for each shift speed. The distribution is made without any variation. In addition, the step ratio value at each gear stage is a value that is far from “1” even at the minimum value thereof, that is, a value that is larger than 1.1 at which a shift feeling can be expected. Therefore, by appropriately allocating the step ratio between the respective gears, a sufficient acceleration feeling can be obtained with a clear shift feeling during a shift involving acceleration.

The invention according to claim 2 is the automatic transmission according to claim 1, further comprising a third control clutch for preventing high speed rotation of the first carrier.
According to the second aspect of the present invention, the same effect as that of the first aspect of the invention can be obtained. In addition, the first pinion can be obtained by disengaging the third control clutch at a predetermined shift. And the situation where the 1st carrier which supports the 1st intermediate pinion reversely rotates at very high speed can be avoided.

  According to a third aspect of the present invention, in the automatic transmission according to the second aspect, the third control clutch selectively connects the input shaft, the first sun gear, and the second carrier. Features.

According to the third aspect of the invention, the same effect as that of the second aspect of the invention can be achieved.
According to a fourth aspect of the present invention, in the automatic transmission according to the second aspect, the third control clutch selectively connects the second ring gear and the third ring gear. .

  According to the fourth aspect of the invention, the same effect as that of the second aspect of the invention can be achieved.

(First embodiment)
Hereinafter, a first embodiment of an automatic transmission according to the present invention will be described with reference to FIGS.

  FIG. 1 is a skeleton diagram showing an automatic transmission 10 according to the present embodiment. The automatic transmission 10 shifts the output rotation of a fluid torque converter 11 that is driven to rotate by, for example, an automobile engine to drive wheels. Used to communicate. As shown in FIG. 1, the automatic transmission 10 includes a transmission case 12 attached to a vehicle body, and a common axis 13 passing through a substantially center in the transmission case 12 from front to rear (from left to right in FIG. 1). ) An input shaft 14 that is sequentially supported, a double planetary gear unit 15 for reduction, a double planetary gear unit 16 for shifting, and an output shaft 17 are provided.

  As shown in FIG. 1, in the double planetary gear unit 15 for speed reduction, a double pinion type first planetary gear mechanism 21 is disposed in the front stage, and a double pinion type second planetary gear mechanism 22 is also provided in the rear stage. It is arranged. Further, in the shift type planetary gear unit 16, the single pinion type third planetary gear mechanism 23 is arranged in the front stage, and the single pinion type fourth planetary gear mechanism 24 is also arranged in the rear stage. .

First, a specific configuration of the reduction type planetary gear unit 15 will be described.
In the reduction type planetary gear unit 15, the first planetary gear mechanism 21 at the front stage includes a first sun gear S 1 rotatably supported on the common axis 13, a first pinion 25 meshing with the first sun gear S 1, and A first intermediate pinion 26 that meshes with the first pinion 25 is provided. Furthermore, the first planetary gear mechanism 21 supports the first pinion 25 and the first intermediate pinion 26 rotatably, and supports the first carrier C1 rotatably supported on the common axis 13 and the first intermediate pinion 26. And a first ring gear R1 that is rotatably supported on the common axis 13.

  On the other hand, the second planetary gear mechanism 22 at the rear stage is connected to the first ring gear R1 and is rotatably supported on the common axis 13, and a second pinion 27 meshing with the second sun gear S0. And a second intermediate pinion 28 that meshes with the second pinion 27. Further, the second planetary gear mechanism 22 meshes with the second carrier C0 and the second intermediate pinion 28 which are supported rotatably on the common axis 13 by supporting the second pinion 27 and the second intermediate pinion 28. The second ring gear R0 is rotatably supported on the common axis 13.

  In the reduction type planetary gear unit 15, the first sun gear S1 and the second carrier C0 are detachably connected to the input shaft 14 by the third control clutch C-3 while being connected to each other. That is, the third control clutch C-3 is provided on a power transmission path capable of transmitting power from the input shaft 14 to the speed-changing double planetary gear device 16 via the speed-reducing double planetary gear device 15. When the third control clutch C-3 is connected, the first sun gear S1 and the second carrier C0 are coupled to the input shaft 14 so that power can be transmitted. The first carrier C1 and the first ring gear R1 are connected to a third control brake B-3 and a first control brake B-1 provided in the transmission case 12, respectively, and the control brakes B-3, B- When 1 operates, each rotation is regulated.

Next, a specific configuration of the shift type planetary gear unit 16 will be described.
In the shift type planetary gear unit 16, the third planetary gear mechanism 23 at the front stage can rotate the third sun gear S <b> 2 that is rotatably supported on the common axis 13 and the third pinion 29 that meshes with the third sun gear S <b> 2. And a third carrier C2 that is rotatably supported on the common axis 13. Further, the third planetary gear mechanism 23 meshes with the third pinion 29 and is connected to the second ring gear R0 of the second planetary gear mechanism 22 in the reduction type double planetary gear device 15 so as to be rotatably supported on the common axis 13. The third ring gear R2 is provided.

  On the other hand, the fourth planetary gear mechanism 24 at the rear stage supports the fourth sun gear S3 rotatably supported on the common axis 13 and the fourth pinion 30 meshing with the fourth sun gear S3 and rotates on the common axis 13. A fourth carrier C3 supported in a possible manner is provided. Further, the fourth planetary gear mechanism 24 meshes with the fourth pinion 30 and is connected to the third carrier C2 of the third planetary gear mechanism 23 at the preceding stage so as to be rotatably supported on the common axis 13. It has.

  In the double planetary gear unit 16 for transmission, the third sun gear S2 and the fourth sun gear S3 are detachably connected to the input shaft 14 by the first control clutch C-1 while being connected to each other. The third carrier C2 and the fourth ring gear R3 are detachably connected to the input shaft 14 by the second control clutch C-2 while being connected to each other. The fourth ring gear R3 is restricted from rotating (reversely rotating) in one direction together with the third carrier C2 of the third planetary gear mechanism 23 in the preceding stage by the one-way clutch F-3 provided in the transmission case 12. A four carrier C3 is connected to the output shaft 17. The third ring gear R2 and the fourth ring gear R3 are respectively connected to a second control brake B-2 and a fourth control brake B-4 provided in the transmission case 12, and the control brakes B-2, B- When 4 operates, each rotation is controlled.

  In the fluid torque converter 11 shown in FIG. 1, the pump impeller 31 is rotationally driven by an engine (not shown) to send out oil, and the stator 32 receives the reaction force of the oil, whereby torque is generated in the turbine 33. It is like that. When the lock-up clutch 34 is operated, the pump impeller 31 and the turbine 33 are directly connected via the lock-up clutch 34, so that torque is also generated in the turbine 33 in this case. Then, by connecting the input shaft 14 to the turbine 33, power is transmitted from the input shaft 14 side to the output shaft 17 via any one of the power transmission paths among the power transmission paths. .

  In the automatic transmission 10 configured as described above, the first to third control clutches C-1 to C-3 and the first to fourth control brakes B-1 to B-4 are selected. The gear ratio of the eight forward speeds and two reverse speeds is controlled by controlling the rotation of each element (sun gear, ring gear, etc.) of the double planetary gear unit 15 for reduction and the double planetary gear unit 16 for transmission. Is established. Therefore, the first to third control clutches C-1 to C-3 and the first to fourth controls at the respective shift speeds (eight forward speed and two reverse speeds) when the automatic transmission 10 is shifted. The operating state of the brakes B-1 to B-4 will be described below with reference to FIG.

  In FIG. 2, the gear ratio (the rotational speed of the input shaft 14 / the rotational speed of the output shaft 17) in each gear stage and the gear when the gear stage is increased by one stage are shown together with the operating state of the control clutch and the like in each gear stage. The step ratio indicating the ratio increase ratio (gear ratio of the current gear / gear ratio of the previous gear) is shown on the right side of the table. In the operation table of FIG. 2, when a white circle is added to each control clutch and control brake column corresponding to each gear, the control clutch is in the connected state, and the control brake is controlled in rotation. It shows that it is in a state. However, as noted below the operation table, a white circle with parentheses indicates that the corresponding control clutch and control brake are in a connection / rotation restricted state during engine braking. Yes. Moreover, when the black circle is attached | subjected, although the applicable control clutch and control brake are engaged, it has shown that it is not concerned in torque transmission (power transmission).

  Here, in each of the double-pinion type planetary gear mechanisms 21 and 22 of the double planetary gear unit 15 for reduction, the rotation speed Ns of the sun gear, the rotation speed Nc of the carrier, the rotation speed Nr of the ring gear, and the planetary gear mechanisms 21 and 22 The relationship with the gear tooth number ratio (number of teeth of the sun gear / number of teeth of the ring gear) λ is expressed by the following equation (1). Further, in the single pinion type planetary gear mechanisms 23 and 24 of the shift type planetary gear device 16, the sun gear rotation speed Ns, the carrier rotation speed Nc, the ring gear rotation speed Nr, and the gears of the planetary gear mechanisms 23 and 24. The relationship with the tooth number ratio λ is expressed by the following equation (2). And the gear ratio in each gear stage is calculated based on this Formula (1), (2).

Nr = (1-λ) Nc + λNs (1)
Nr = (1 + λ) Nc−λNs (2)
If the number of teeth of the sun gears S1, S0, S2, S3 is Zs1, Zs0, Zs2, Zs3, and the number of teeth of the ring gears R1, R0, R2, R3 is Zr1, Zr0, Zr2, Zr3, a double planetary gear unit for reduction is used. The gear gear ratios of the planetary gear mechanisms 21 and 22 of 15 and the planetary gear mechanisms 23 and 24 of the double planetary gear device 16 for shifting are λ1 = Zs1 / Zr1, λ0 = Zs0 / Zr0, λ2 = Zs2 / Zr2. , Λ3 = Zs3 / Zr3. The gear tooth number ratios λ1, λ0, λ2, and λ3 of the planetary gear mechanisms 21 to 24 thus obtained are shown on the upper side of the table in FIG.

  In such an automatic transmission 10, the control clutches C-1 to C-3 and the control brakes B-1 to B-4 are selectively engaged and disengaged as shown in the operation table of FIG. In this case, the speed ratio of the elements (sun gear, ring gear, etc.) of the planetary gear mechanisms 21 to 24 in the planetary gear devices 15 and 16 is as shown in the velocity diagram shown in FIG. That is, in this velocity diagram, the elements including the sun gears S0 to S3, the carriers C0 to C3, and the ring gears R0 to R3 of the planetary gear units 15 and 16 are associated with the gear tooth number ratios λ0 to λ3 in the horizontal axis direction. They are arranged at intervals, and the speed ratio is taken corresponding to each element in the vertical axis direction. In the velocity diagram of FIG. 3, the velocity diagrams of the speed reduction double planetary gear device 15 and the speed change double planetary gear device 16 are shown side by side.

  First, in the speed diagram of the left-side compound planetary gear unit 15 for reduction, the first ring gear R1 and the second sun gear S0, and the first sun gear S1 and the second carrier C0 are connected to each other and are common, so R1, S0. In addition, the respective speed ratios of the first ring gear R1 and the second sun gear S0, and the first sun gear S1 and the second carrier C0 are displayed on one vertical line to which S1 and S0 are respectively attached. Further, the respective speed ratios of the first carrier C1 and the second ring gear R0 are displayed on one vertical line to which C1 and R0 are respectively attached. In both the first and second planetary gear mechanisms 21 and 22 of the double pinion type, the intervals between the vertical lines of the carriers C1 and C0 and the vertical lines of the sun gears S1 and S0 are regarded as “1”, respectively. The vertical lines of the ring gears R1 and R0 are arranged on the same side as the vertical lines of the sun gears S1 and S0 from the vertical lines of the carriers C1 and C0 by an interval corresponding to the gear tooth ratios λ0 and λ1, respectively. .

  On the other hand, since the fourth ring gear R3 and the third carrier C2 and the fourth sun gear S3 and the third sun gear S2 are connected to each other and are common in the speed diagram of the right-side double planetary gear device 16 for transmission, R3, C2 In addition, the respective speed ratios of the fourth ring gear R3 and the third carrier C2, and the fourth sun gear S3 and the third sun gear S2 are displayed on one vertical line to which S3 and S2 are respectively attached. Further, the respective speed ratios of the third ring gear R2 and the fourth carrier C3 are displayed on one vertical line to which R2 and C3 are respectively attached. In both the single-pinion type third planetary gear mechanism 23 and the fourth planetary gear mechanism 24, the intervals between the vertical lines of the carriers C2 and C3 and the vertical lines of the sun gears S2 and S3 are regarded as “1”, respectively. The vertical lines of the ring gears R2 and R3 are spaced apart from the vertical lines of the carriers C2 and C3 on the opposite side of the vertical lines of the sun gears S2 and S3 by an interval corresponding to the gear tooth ratios λ2 and λ3.

  In the speed diagram of FIG. 3, the first to fourth control brakes B-1 to B-4 and the first to third control clutches C-1 to C-3 are selectively operated. The symbols B-1 to B-4 and C-1 to C-3 are entered. In addition, elements corresponding to each other when power is transmitted at each gear position between the speed diagram of the left-side double planetary gear unit 15 for speed reduction and the speed diagram of the right-side compound planetary gear unit 16 for shifting. By displaying the connection with broken lines, the power transmission path at each gear stage is indicated.

  In the speed diagram of the right-side compound planetary gear unit 16 for shifting, the elements corresponding to the four vertical lines are the first, second, third, and fourth elements in the order of the vertical lines. The third ring gear R2 as the first element is connected to the second ring gear R0 of the double planetary gear unit 15 for reduction, and the fourth ring gear R3 and the third carrier C2 connected as the second element are connected to the one-way clutch F. -3 is connected in parallel to the second control clutch C-2 and the fourth control brake B-4 in a state where rotation (reverse rotation) in one direction is restricted. The fourth carrier C3 as the third element is connected to the output shaft 17, and the fourth sun gear S3 and the third sun gear S2 as the fourth element are connected to each other by the first control clutch C-1. The shaft 14 is detachably connected.

Therefore, next, the operation of each gear stage in the automatic transmission 10 configured as described above will be described with reference to FIG.
First, in the case of the first forward speed, the third sun gear S2 and the fourth sun gear S3 are connected to the input shaft 14 by the operation of the first control clutch C-1, and the input shaft is connected to the third sun gear S2 and the fourth sun gear S3. 14 rotations are transmitted. In this case, since the fourth ring gear R3 is restricted from being driven in reverse by the operation of the one-way clutch F-3, the fourth pinion 30 meshing with the fourth sun gear S3 is moved to the fourth ring gear R3 in which the reverse drive is restricted. The fourth carrier C3 as the third element that supports the fourth pinion 30 rotates by revolving with the reaction force supported. As a result, the output shaft 17 connected to the fourth carrier C3 is driven to rotate forward at a gear ratio 3.5385 of the first forward shift speed shown in FIG. During engine braking, the one-way clutch F-3 idles and the reverse rotation of the fourth ring gear R3 cannot be regulated. In this case, the fourth control brake B-4 is activated to activate the fourth ring gear R3. By restricting the rotation, the fourth carrier C3 and the output shaft 17 are allowed to rotate while allowing the fourth pinion 30 to revolve.

  Next, in the case of the second forward speed, the third sun gear S2 and the fourth sun gear S3 are connected to the input shaft 14 by the operation of the first control clutch C-1, and input to the third sun gear S2 and the fourth sun gear S3. The rotation of the shaft 14 is transmitted. In this case, since the rotation of the third ring gear R2 is restricted by the operation of the second control brake B-2, the third pinion 29 meshing with the third sun gear S2 revolves, and the third carrier C2 and the fourth carrier The ring gear R3 is rotated. Then, the fourth pinion 30 revolves in accordance with the rotational difference between the fourth ring gear R3 and the fourth sun gear S3, and the fourth carrier C3 as the third element that supports the fourth pinion 30 rotates. The output shaft 17 connected to the carrier C3 is driven to rotate forward at a gear ratio 2.0604 of the second forward shift speed shown in FIG.

  Next, in the case of the third forward speed, the third sun gear S2 and the fourth sun gear S3 are connected to the input shaft 14 by the operation of the first control clutch C-1, and input to the third sun gear S2 and the fourth sun gear S3. The rotation of the shaft 14 is transmitted. Further, the first sun gear S1 and the second carrier C0 are connected to the input shaft 14 by the operation of the third control clutch C-3, and the rotation of the input shaft 14 is also transmitted to the first sun gear S1 and the second carrier C0. In this case, the first ring gear R1 and the second sun gear S0 coupled to each other are restricted by the operation of the first control brake B-1, so that the second sun gear is rotated along with the rotation of the second carrier C0. The second pinion 27 meshing with S0 and the second intermediate pinion 28 meshing with the second pinion 27 revolve.

  Then, the second ring gear R0 meshing with the second intermediate pinion 28 rotates together with the third ring gear R2 connected to each other, and the third pinion 29 is changed according to the rotational difference between the third ring gear R2 and the third sun gear S2. Revolves and rotates the third carrier C2 and the fourth ring gear R3. As a result of the fourth pinion 30 revolving according to the rotational difference between the fourth ring gear R3 and the fourth sun gear S3, the fourth carrier C3 as the third element supporting the fourth pinion 30 is rotated. The output shaft 17 connected to the carrier C3 is driven to rotate forward at a gear ratio of 1.3281 of the third forward speed shown in FIG.

  Next, in the case of the fourth forward speed, the third sun gear S2 and the fourth sun gear S3 are connected to the input shaft 14 by the operation of the first control clutch C-1, and input to the third sun gear S2 and the fourth sun gear S3. The rotation of the shaft 14 is transmitted. Further, the first sun gear S1 and the second carrier C0 are connected to the input shaft 14 by the operation of the third control clutch C-3, and the rotation of the input shaft 14 is also transmitted to the first sun gear S1 and the second carrier C0. In this case, since the rotation of the first carrier C1 is restricted by the operation of the third control brake B-3, the first intermediate pinion 26 supported by the first carrier C1 is rotated with the rotation of the first sun gear S1. The meshing first ring gear R1 rotates together with the second sun gear S0 connected to each other.

  Then, with the rotation of the second sun gear S0, the second ring gear R0 meshing with the second intermediate pinion 28 supported by the second carrier C0 rotates together with the third ring gear R2 connected to each other, and this third ring gear. The third pinion 29 revolves according to the rotational difference between R2 and the third sun gear S2. When the third carrier C2 and the fourth ring gear R3 rotate along with the revolution of the third pinion 29, the fourth pinion 30 revolves according to the rotational difference between the fourth ring gear R3 and the fourth sun gear S3, The 4th carrier C3 as a 3rd element which supports 4 pinion 30 rotates. As a result, the output shaft 17 connected to the fourth carrier C3 is driven to rotate forward at the gear ratio 1.1605 of the fourth forward speed shown in FIG.

  Next, in the case of the fifth forward speed, the third sun gear S2 and the fourth sun gear S3 are connected to the input shaft 14 by the operation of the first control clutch C-1, and input to the third sun gear S2 and the fourth sun gear S3. The rotation of the shaft 14 is transmitted. Further, the third carrier C2 and the fourth ring gear R3 coupled to each other are connected to the input shaft 14 by the operation of the second control clutch C-2, and the rotation of the input shaft 14 is also performed to the third carrier C2 and the fourth ring gear R3. Is transmitted. As a result, the fourth carrier C3 as a third element for supporting the fourth pinion 30 meshing with the fourth sun gear S3 and the fourth ring gear R3 also rotates together, and the output shaft connected to the fourth carrier C3. 17 is driven to rotate forward at a gear ratio of 1.000 of the fifth forward speed shown in FIG.

  Next, in the case of the sixth forward speed, the third carrier C2 and the fourth ring gear R3 connected to each other are connected to the input shaft 14 by the operation of the second control clutch C-2, and the third carrier C2 The rotation of the input shaft 14 is transmitted to the 4-ring gear R3. Further, the first sun gear S1 and the second carrier C0 are connected to the input shaft 14 by the operation of the third control clutch C-3, and the rotation of the input shaft 14 is also transmitted to the first sun gear S1 and the second carrier C0. In this case, since the rotation of the first carrier C1 is restricted by the operation of the third control brake B-3, the first intermediate pinion 26 supported by the first carrier C1 is rotated with the rotation of the first sun gear S1. The meshing first ring gear R1 rotates together with the second sun gear S0 connected to each other.

  Then, the second ring gear R0 rotates together with the third ring gear R2 connected to each other according to the rotational difference between the second sun gear S0 and the second pinion 27 and the rotational difference between the second pinion 27 and the second intermediate pinion 28. The third sun gear S2 rotates with the fourth sun gear S3 connected to each other in accordance with the rotational difference between the third ring gear R2 and the third carrier C2. Then, the fourth pinion 30 revolves according to the rotational difference between the fourth sun gear S3 and the fourth ring gear R3, and the fourth carrier C3 as a third element that supports the fourth pinion 30 rotates. As a result, the output shaft 17 connected to the fourth carrier C3 is driven to rotate forward at a gear ratio 0.8383 of the sixth forward speed shown in FIG.

  Next, in the case of the seventh forward shift speed, the third carrier C2 and the fourth ring gear R3 coupled to each other are connected to the input shaft 14 by the operation of the second control clutch C-2, and the third carrier C2 The rotation of the input shaft 14 is transmitted to the 4-ring gear R3. Further, the first sun gear S1 and the second carrier C0 are connected to the input shaft 14 by the operation of the third control clutch C-3, and the rotation of the input shaft 14 is also transmitted to the first sun gear S1 and the second carrier C0. In this case, the first ring gear R1 and the second sun gear S0 coupled to each other are restricted by the operation of the first control brake B-1, so that the second sun gear is rotated along with the rotation of the second carrier C0. The second pinion 27 meshing with S0 and the second intermediate pinion 28 meshing with the second pinion 27 revolve.

  Then, the second ring gear R0 meshing with the second intermediate pinion 28 rotates with the third ring gear R2 connected to each other, and the third sun gear S2 is rotated according to the rotational difference between the third ring gear R2 and the third carrier C2. It rotates with the 4th sun gear S3 connected mutually. Then, the fourth pinion 30 revolves according to the rotational difference between the fourth sun gear S3 and the fourth ring gear R3, and the fourth carrier C3 as a third element that supports the fourth pinion 30 rotates. As a result, the output shaft 17 connected to the fourth carrier C3 is driven to rotate forward at a gear ratio 0.7439 of the seventh forward speed shown in FIG.

  Next, in the case of the eighth forward shift speed, the third carrier C2 and the fourth ring gear R3 coupled to each other are connected to the input shaft 14 by the operation of the second control clutch C-2, and the third carrier C2 The rotation of the input shaft 14 is transmitted to the 4-ring gear R3. In this case, since the rotation of the third ring gear R2 is restricted by the operation of the second control brake B-2, the third sun gear S2 rotates with the fourth sun gear S3 connected to each other, and the fourth sun gear. The fourth pinion 30 revolves according to the rotational difference between S3 and the fourth ring gear R3, and the fourth carrier C3 as a third element that supports the fourth pinion 30 rotates. As a result, the output shaft 17 connected to the fourth carrier C3 is driven forward at a gear ratio of 0.5823 of the eighth forward shift speed shown in FIG.

  Next, in the case of the reverse first speed, the first sun gear S1 and the second carrier C0 are connected to the input shaft 14 by the operation of the third control clutch C-3, and input to the first sun gear S1 and the second carrier C0. The rotation of the shaft 14 is transmitted. In this case, the first ring gear R1 and the second sun gear S0 coupled to each other are restricted by the operation of the first control brake B-1, so that the second sun gear is rotated along with the rotation of the second carrier C0. The second pinion 27 meshing with S0 and the second intermediate pinion 28 meshing with the second pinion 27 revolve. As this revolution occurs, the second ring gear R0 meshing with the second intermediate pinion 28 rotates together with the third ring gear R2 connected to each other. In this case, since the fourth ring gear R3 and the third carrier C2 coupled to each other are restricted by the operation of the fourth control brake B-4, the third ring C supported by the third carrier C2 is restricted. The third sun gear S2 rotates in reverse with the fourth sun gear S3 connected to each other via the pinion 29. Then, the fourth pinion 30 meshing with the fourth sun gear S3 revolves with the reaction force supported by the fourth ring gear R3, and the fourth carrier C3 as the third element that supports the fourth pinion 30 rotates. As a result, the output shaft 17 connected to the fourth carrier C3 is reversely driven at a predetermined gear ratio of the reverse first gear.

  Next, in the case of the second reverse speed, the first sun gear S1 and the second carrier C0 are connected to the input shaft 14 by the operation of the third control clutch C-3, and input to the first sun gear S1 and the second carrier C0. The rotation of the shaft 14 is transmitted. In this case, since the rotation of the first carrier C1 is restricted by the operation of the third control brake B-3, the first intermediate pinion 26 supported by the first carrier C1 is rotated with the rotation of the first sun gear S1. The meshed first ring gear R1 rotates with the second sun gear S0 connected to each other, and the second ring gear R0 rotates with the third ring gear R2 in accordance with the turning point difference between the second sun gear S0 and the second carrier C0.

  In this case, since the fourth ring gear R3 and the third carrier C2 coupled to each other are restricted by the operation of the fourth control brake B-4, the third ring C supported by the third carrier C2 is restricted. The third sun gear S2 rotates in reverse with the fourth sun gear S3 connected to each other via the pinion 29. Then, the fourth pinion 30 meshing with the fourth sun gear S3 revolves with the reaction force supported by the fourth ring gear R3, and the fourth carrier C3 as the third element that supports the fourth pinion 30 rotates. As a result, the output shaft 17 connected to the fourth carrier C3 is reversely driven at a predetermined gear ratio of the reverse second speed.

  In the above-described first forward shift speed, the third ring gear R2 rotates in reverse according to the rotation of the third sun gear S2, but the second ring gear R0 interconnected with the third ring gear R2 also rotates in reverse. Therefore, when the third control clutch C-3 is not provided, the rotation of the input shaft 14 is also transmitted to the first sun gear S1 and the second carrier C0, and the rotation of the second carrier C0 and the second ring gear R0. Depending on the difference, the second sun gear S0 rotates with the first ring gear R1 connected to each other. Then, as a result of a large relative rotational difference between the first ring gear R1 and the first sun gear S1, the first pinion 25 and the first intermediate pinion 26 that mesh with the first ring gear R1 and the first sun gear S1 are supported. One carrier C1 rotates at a very high speed. However, in the case of the automatic transmission 10 of the present embodiment, the third control clutch C-3 is provided, and the third control clutch C-3 is disconnected at the first forward shift speed. A very high speed rotation of the first carrier C1 as described above is avoided.

  In the automatic transmission 10 of the present embodiment, each gear stage is in the operating state as described above at the time of shifting, and each sun gear S0 to S3 and each carrier C0 to C3 in each gear stage when the rotational speed of the input shaft 14 is 1. , And the rotation ratios of the ring gears R0 to R3 are as shown in the velocity diagram of FIG. Therefore, as is apparent from the speed diagram of FIG. 3, the rotation ratio, that is, the gear ratio of the fourth carrier C3, which is the third element in each shift stage, is arranged at an appropriate interval without significant variation, and is appropriately separated. A gear ratio of 8 forward speeds and 2 reverse speeds can be realized.

  Further, as shown in FIG. 2, the step ratio, which is the rate of increase of the gear ratio when the gear stage is increased by one stage, is 1.717 between the first and second gear stages, and the second and third gear stages. Between the first and second gears is 1.144, between the fourth and fifth gears is 1.161, between the fifth and sixth gears is 1.193, 6 between the seventh and seventh gears and 1.277 between the seventh and eighth gears. That is, this step ratio is also distributed with no great variation for each gear position. The value of the step ratio at each gear is 1.127 even in the step ratio between the sixth and seventh gears, which is the minimum value among the step ratios.

Therefore, according to the automatic transmission 10 of the present embodiment, the following effects can be obtained.
(1) The step ratio in each of the eight forward speeds is distributed in a state in which there is no large variation among the respective speeds. Also, the step ratio value of each gear stage is 1.127 even at the step ratio between the sixth and seventh gear speeds, which is the minimum value thereof, and a value away from “1”, that is, a shift feeling is brought about. This is a value larger than 1.1 that can be expected. Therefore, by appropriately allocating the step ratio between the respective gears, a sufficient acceleration feeling can be obtained with a clear shift feeling during a shift involving acceleration.

  (2) Further, at the time of shifting at the first forward speed, the third control clutch C-3 is disengaged, and the first sun gear S1 and the first ring gear R1 rotate with a large relative rotational difference. Therefore, it is possible to avoid a situation in which the first carrier C1 supporting the first pinion 25 and the first intermediate pinion 26 reversely rotates at a very high speed.

  (3) Further, the third control clutch C-3 can be arranged in the transmission case 12 closer to the front than the place where the planetary gear units 15 and 16 are arranged. Therefore, it is possible to form an oil passage in the input shaft 14 along the common axis 13 and supply hydraulic oil to the third control clutch C-3 through this oil passage, with respect to the third control clutch C-3. It is easy to secure an oil passage for supplying hydraulic oil.

(Second Embodiment)
Next, a second embodiment of the automatic transmission according to the present invention will be described with reference to FIGS. The second embodiment is different from the first embodiment in the location of the third control clutch, and the other configurations are the same as those in the first embodiment. Therefore, in the following description, parts different from those of the first embodiment will be mainly described, and the same reference numerals are given to common member configurations and the like, and redundant description will be omitted.

  In the automatic transmission 10 of the present embodiment, as shown in FIG. 4, the first sun gear S1 and the second planetary gear mechanism of the first planetary gear mechanism 21 connected to each other in the double planetary gear unit 15 for reduction. The 22 second carriers C0 are connected to the input shaft 14, and the rotation of the input shaft 14 is transmitted to the first sun gear S1 and the second carrier C0. On the other hand, the second ring gear R0 of the second planetary gear mechanism 22 in the double planetary gear unit 15 for reduction and the third ring gear R2 of the third planetary gear mechanism 23 in the double planetary gear unit 16 for shifting are the third control clutch C−. 3 is detachably connected. In other respects, as shown in FIG. 4, each member configuration in the automatic transmission 10 is the same as that in the first embodiment.

  Also in the second embodiment, the operation state at the time of shifting of each gear stage is as shown in the operation table of FIG. 2 as in the case of the first embodiment. That is, at the third forward speed, the fourth forward speed, the sixth forward speed, and the seventh forward speed, the third control clutch C-3 is operated. As a result, the second ring gear R0 and the third ring gear are operated. R2 comes to be connected. The third control clutch C-3 is engaged even at the fifth shift speed, but is not related to torque (power) transmission.

  Further, the rotation ratios of the sun gears S0 to S3, the carriers C0 to C3, and the ring gears R0 to R3 at the respective speeds when the rotation speed of the input shaft 14 is 1 are as shown in the velocity diagram of FIG. become. Therefore, also in the second embodiment, the eight forward gears that are arranged at appropriate intervals without causing a large variation in the rotation ratio, that is, the gear ratio, of the fourth carrier C3 that is the third element in each gear, A reverse gear ratio of two stages can be realized. Further, the step ratio, which is the rate of increase of the gear ratio when the gear position is increased by one, is the respective step ratios as shown in FIG. 2 as in the first embodiment.

Therefore, also in the automatic transmission 10 of the second embodiment, the same operational effects as the above (1) and (2) in the first embodiment can be achieved.
(Third embodiment)
Next, a third embodiment according to the automatic transmission of the present invention will be described with reference to FIGS. The third embodiment is different from the first embodiment in that it does not have a third control clutch, and the other configuration is the same as that of the first embodiment. Therefore, in the following description, parts different from those of the first embodiment will be mainly described, and the same reference numerals are given to common member configurations and the like, and redundant description will be omitted.

  Now, in the automatic transmission 10 of the present embodiment, as shown in FIG. 6, the first sun gear S1 and the second planetary gear mechanism of the first planetary gear mechanism 21 connected to each other in the double planetary gear unit 15 for reduction. The 22 second carriers C0 are connected to the input shaft 14, and the rotation of the input shaft 14 is transmitted to the first sun gear S1 and the second carrier C0. In other respects, as shown in FIG. 6, each member configuration in the automatic transmission 10 is the same as that in the first embodiment.

  And in this 3rd Embodiment, when the operation state at the time of the shift of each gear stage is contrasted with 1st Embodiment, since it does not have the 3rd control clutch C-3, it is forward 1st. Unlike the case of the first embodiment, the first carrier C1 rotates at the shift speed, the second forward speed, the eighth forward speed, the first reverse speed, and the second reverse speed. .

  That is, in each of these shift speeds, the third ring gear R2 rotates in reverse according to the rotation of the third sun gear S2, but the second ring gear R0 interconnected with the third ring gear R2 also rotates in reverse. Therefore, when the third control clutch C-3 is not provided, the rotation of the input shaft 14 is also transmitted to the first sun gear S1 and the second carrier C0, and the rotation of the second carrier C0 and the second ring gear R0. Depending on the difference, the second sun gear S0 rotates with the first ring gear R1 connected to each other. As a result, the first carrier C1 that supports the first pinion 25 and the first intermediate pinion 26 that mesh with the first ring gear R1 and the first sun gear S1, respectively, rotates. In this case, the rotational speed of the first carrier C1 corresponds to the magnitude of the relative rotational difference between the first ring gear R1 and the first sun gear S1. Therefore, in the case of the first forward shift speed, it rotates at the highest speed.

  Further, the rotation ratios of the sun gears S0 to S3, the carriers C0 to C3, and the ring gears R0 to R3 at the respective speeds when the rotation speed of the input shaft 14 is 1 are as shown in the velocity diagram of FIG. become. Therefore, also in the third embodiment, the eight forward gears that are arranged at appropriate intervals without causing a large variation in the rotation ratio, that is, the gear ratio, of the fourth carrier C3 that is the third element in each gear, A reverse gear ratio of two stages can be realized. Further, the step ratio, which is the rate of increase of the gear ratio when the shift speed is increased by one, is the respective step ratios as shown in FIG. 7, as in the case of the first embodiment.

Therefore, also in the automatic transmission 10 of the third embodiment, it is possible to achieve the same function and effect as the above (1) in the first embodiment.
In addition, you may change said each embodiment into another embodiment (another example) as follows.

  In each of the above embodiments, if the gear tooth ratios λ0, λ1, λ2, and λ3 shown in the operation tables of FIGS. 2 and 7 are satisfied, the sun gears S0 to S3 in the planetary gear mechanisms 21 to 24 The number of teeth of each ring gear R0-R3 can be set arbitrarily.

  -In the said 2nd Embodiment, if the 3rd control clutch C-3 can connect between 2nd ring gear R0 and 3rd ring gear R2 so that engagement / disengagement is possible, the specific arrangement | positioning location is arbitrary. is there.

  In each of the first and second embodiments, the engagement relationship at each gear stage indicated by the black circles of the third control clutch C-3 and the first control brake B-1 in the operation table of FIG. It may be.

The skeleton figure of the automatic transmission of 1st Embodiment. The operation | movement table | surface of the control clutch and control brake in each gear stage similarly. The speed diagram which similarly shows the gear ratio of each element of the planetary gear apparatus in each gear stage. The skeleton figure of the automatic transmission of 2nd Embodiment. The speed diagram which similarly shows the gear ratio of each element of the planetary gear apparatus in each gear stage. The skeleton figure of the automatic transmission of 3rd Embodiment. The operation | movement table | surface of the control clutch and control brake in each gear stage similarly. The speed diagram which similarly shows the gear ratio of each element of the planetary gear apparatus in each gear stage.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 10 ... Automatic transmission, 14 ... Input shaft, 15 ... Double planetary gear apparatus for reduction, 16 ... Double planetary gear apparatus for transmission, 17 ... Output shaft, 21-24 ... Planetary gear mechanism, 25, 27, 29, 30 ... Pinion, 26, 28 ... Intermediate pinion, S0, S1, S2, S3 ... Sun gear, C0, C1, C2, C3 ... Carrier, R0, R1. R2, R3 ... ring gear, C-1 to C-3 ... control clutch, B-1 to B-4 ... control brake, [lambda], [lambda] 0, [lambda] 1, [lambda] 2, [lambda] 3 ... gear tooth ratio.

Claims (4)

A speed reduction compound planetary gear unit having both a double-pinion type first planetary gear mechanism and a second planetary gear mechanism, and a gear-shifting compound planetary gear unit having both a single-pinion type third planetary gear mechanism and a fourth planetary gear mechanism. Prepared,
In the double planetary gear unit for reduction, the first carrier that supports the first planetary gear mechanism on a first sun gear, a first pinion that meshes with the first sun gear, and a first intermediate pinion that meshes with the first pinion. A first ring gear meshing with the first intermediate pinion, a second sun gear coupled to the first ring gear, and a second pinion meshing with the second sun gear. A second carrier that supports the second intermediate pinion that meshes with the second pinion; a second ring gear that meshes with the second intermediate pinion; and the first sun gear and the second carrier connected to each other. The first carrier is connected to the third control brake and the first ring gear and the second sun gear are connected to the first control brake. And,
In the double planetary gear device for speed change, the third planetary gear mechanism is engaged with the third sun gear, a third carrier that supports the third pinion that meshes with the third gear, and the second ring gear that meshes with the third pinion. And a third ring gear coupled to be able to transmit power to the fourth planetary gear mechanism. The fourth planetary gear mechanism includes a fourth sun gear, a fourth carrier that supports a fourth pinion that meshes with the fourth sun gear, and the fourth carrier. A fourth ring gear meshing with a pinion and coupled to the three carriers is provided, and the third sun gear and the fourth sun gear are coupled to the input shaft so as to be disengageable by a first control clutch. In addition, the third ring gear and the fourth ring gear are connected to the second control brake and the fourth control brake, respectively, and the third carrier is engaged with the input shaft by the second control clutch. While linked, coupled to an output shaft of the fourth carrier, the automatic transmission, characterized in that rotation of the second ring gear is transmitted to the third ring gear. The automatic transmission according to claim 1, further comprising a third control clutch for preventing the first carrier from rotating at a high speed. The automatic transmission according to claim 2, wherein the third control clutch selectively connects the input shaft, the first sun gear, and the second carrier. The automatic transmission according to claim 2, wherein the third control clutch selectively connects the second ring gear and the third ring gear.

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