JP2016044727A - Non-stage transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a non-stage transmission having strength which matches with torque fluctuations associated with the operation strokes of an internal combustion engine, and capable of achieving downsizing and lightening or cost reduction.SOLUTION: A non-stage transmission comprises: an input shaft to which driving force is transmitted from an internal combustion engine having plural cylinders; an output shaft; plural lever crank mechanisms each having an eccentric mechanism which has a cam disc eccentric to the input shaft and a rotary disc rotating eccentrically to the cam disc, and a swing link pivotally supported on the output shaft, and converting a rotary motion of the input shaft to swing motions of the swing links; and a mechanism for preventing unidirectional rotation of the swing links relative to the output shaft. The cam discs of the respective eccentric mechanisms are arranged eccentrically at every predetermined angle in the circumferential direction of the input shaft, and when setting a rotation angle position of the input shaft at the time point of passing over the top dead center in a stroke from a compression stroke to an expansion stroke in each cylinder as a reference position, a rotation angle difference between the reference positions of the respective cylinders whose orders of combustion are adjacent is different from a predetermined angle. At least one of the lever crank mechanisms is different from the other in strength against the driving force from the internal combustion engine.SELECTED DRAWING: Figure 9

Description

  The present invention relates to a crank type continuously variable transmission.

  Patent Document 1 discloses a plurality of crank-type transmission units that convert rotation of an input shaft connected to an engine into reciprocating motion of a connecting rod, and convert reciprocating motion of the connecting rod into rotational motion of an output shaft by a one-way clutch. A continuously variable transmission provided and a control device for the continuously variable transmission are disclosed.

  Each transmission unit of the continuously variable transmission disclosed in Patent Document 1 includes a fixed disk that is eccentrically provided on an input shaft, and a swinging disk that is eccentrically provided on the fixed disk and is rotatably provided. A one-way clutch is provided between the swing link and the output shaft. The one-way clutch fixes the swing link to the output shaft when the swing link is about to rotate relative to the output shaft, and To idle the swing link.

  A pinion shaft is inserted into the input shaft, and a notch hole is formed at a location facing the eccentric direction of the fixed disk, and the pinion shaft is exposed from the notch hole. The swing disk is provided with a receiving hole for receiving the input shaft and the fixed disk. Inner teeth are formed on the inner peripheral surface of the swing disk that forms the receiving hole. The inner teeth mesh with the pinion shaft exposed from the notch hole of the input shaft. When the input shaft and the pinion shaft are rotated at the same speed, the eccentric amount of the eccentric mechanism in the transmission unit is maintained. When the rotational speeds of the input shaft and the pinion shaft are made different, the amount of eccentricity of the eccentric mechanism in the transmission unit is changed, and the gear ratio changes.

  When the eccentric mechanism of the transmission unit is rotated by rotating the input shaft, the large-diameter annular portion of the connecting rod rotates, and the swing end of the swing link connected to the other end of the connecting rod Swing. Since the swing link is provided on the output shaft via the one-way clutch, the rotational drive force (torque) is transmitted to the output shaft only when rotating to one side.

JP 2012-251608 A

  A fatigue strength design corresponding to the maximum load applied to the input shaft is uniformly applied to each transmission unit of the continuously variable transmission described above. However, depending on the number of cylinders of the engine and the number of transmission units of the continuously variable transmission, the load applied to the connecting rod for transmitting torque may vary depending on the transmission unit.

In a series of operation strokes of the engine, the reciprocating motion of the piston is taken out as a rotational driving force with torque fluctuation by the crankshaft through the connecting rod. The load accompanying this rotational driving force is applied from the input shaft of the continuously variable transmission to the connecting rod of the transmission unit. In order for the load accompanying the rotational driving force of the engine to be equally applied to each connecting rod of the continuously variable transmission, the following conditions (1) to (2) must be satisfied.
(1) The relationship between the number of cylinders n of the engine and the number m of the transmission units of the continuously variable transmission is n = α · m (where n, m = natural number, n, m ≠ 1).
(2) The fixed disk is eccentrically arranged on the input shaft of the continuously variable transmission at every angle that matches the crank angle (θc / n) obtained by dividing the crank angle θc required for the entire stroke of the engine by the number of cylinders n. Has been.

  Specifically, in each cylinder of an engine of four cycles (intake stroke, compression stroke, expansion stroke (combustion stroke), exhaust stroke), a series of operation strokes is completed at a crank angle of 720 degrees. When applied to an engine, each cylinder is designed to reach an expansion stroke at a crank angle of 180 degrees. Since the torque output in the expansion stroke in each cylinder is relatively larger than the torque in the other strokes, the output torque of the engine as a whole pulsates. For this reason, when the continuously variable transmission includes, for example, two transmission units, and the fixed disk is eccentrically arranged at every 180 degrees with respect to the input shaft, the cycle at which the connecting rod of each transmission unit receives torque is in each cylinder of the engine. Since it overlaps with the timing of the expansion stroke, the connecting rod of either transmission unit of the continuously variable transmission receives the same torque.

  On the other hand, when the continuously variable transmission includes, for example, three transmission units and the fixed disk is eccentrically arranged with respect to the input shaft every 120 degrees, the connecting rod of each transmission unit receives the torque and each of the four cycle engines. It is different from the timing of the expansion stroke in the cylinder. For this reason, conventionally, the design is uniformly applied to all the transmission units in the continuously variable transmission based on the fatigue strength design for the transmission unit to which the largest torque is applied.

  Generally, in order to increase the strength against torque, the member is made of a material having a high rigidity by increasing the cross-sectional area of the member with respect to the torque transmission path. As a result, the weight of the member may increase, the shape may increase, and the cost may increase. For this reason, in the above-described continuously variable transmission, the fatigue strength design is appropriate for the transmission unit to which the largest torque is applied, but the fatigue strength design is not only over-spec for the transmission units other than the transmission unit. This hinders the reduction in size and weight and cost.

  An object of the present invention is to provide a continuously variable transmission that can be reduced in size, weight, and cost, having strength according to torque fluctuations accompanying an operation stroke of an internal combustion engine.

In order to achieve the above object, the invention described in claim 1
A driving force is transmitted from an internal combustion engine (for example, an engine E in an embodiment described later) having a plurality of cylinders (for example, cylinders C1, C2, C3, C4 in an embodiment described later) mounted on the vehicle. An input shaft (for example, the input shaft 2 in the embodiment described later);
An output shaft (for example, an output shaft 3 in an embodiment described later) disposed in parallel with the input shaft;
A cam disk (e.g., a cam disk 5 in an embodiment described later) provided eccentrically with the input shaft and a rotary disk (e.g., a rotation in an embodiment described later) provided eccentrically with respect to the cam disk. An eccentric mechanism (for example, a turning radius adjusting mechanism 4 in an embodiment described later) having a disk 6), and a swing link (for example, a swing link 18 in an embodiment described later) supported by the output shaft; A plurality of lever crank mechanisms (for example, a lever crank mechanism 20 in an embodiment described later) that converts the rotational motion of the input shaft into the swing motion of the swing link;
The swing link is fixed to the output shaft when the swing link is about to rotate to one side around the output shaft, and the swing shaft is fixed to the output shaft when the swing link is about to rotate to the other side. A continuously variable transmission (for example, a continuously variable transmission 1 in an embodiment described later) including a one-way rotation prevention mechanism (for example, a one-way clutch 17 in an embodiment described later) that idles the swing link. And
The cam disks in the turning radius adjusting mechanisms of the plurality of lever crank mechanisms are arranged eccentrically at predetermined angles in the circumferential direction of the input shaft, and in the stroke from the compression stroke to the expansion stroke in each cylinder of the internal combustion engine. When the rotation angle position of the input shaft at the time when the top dead center is exceeded is set as a reference position, the rotation angle difference between the reference positions of adjacent cylinders in the combustion order is different from a predetermined angle, At least one of the lever crank mechanisms is different in strength against the driving force from the internal combustion engine from other lever crank mechanisms.

The invention according to claim 2 is the invention according to claim 1,
Among the plurality of lever crank mechanisms, the lever crank mechanism having the smallest relative angle between the rotation angle of the input shaft at the reference position and the eccentric angle of the cam disk with respect to the input shaft is driven from the internal combustion engine. Strength against force is higher than other lever crank mechanisms.

The invention according to claim 3 is the invention according to claim 1 or 2,
The lever crank mechanism has a large-diameter annular portion (for example, a large-diameter annular portion 15a in an embodiment described later) that is rotatably fitted to the eccentric mechanism at one end portion, and the other end portion is A connecting rod (for example, a connecting rod 15 in an embodiment described later) connected to the swing end of the swing link;
The width of the connecting rod of the lever crank mechanism with high strength is wider than the width of the connecting rod of the lever crank mechanism with low strength.

The invention according to claim 4 is the invention according to any one of claims 1 to 3,
The reference position is the rotational angle position of the input shaft after the expansion stroke starts, and the torque received by the input shaft at the reference position is the maximum torque obtained in the stroke of each cylinder of the internal combustion engine. is there.

The invention according to claim 5 is the invention according to any one of claims 1 to 4,
The number of the plurality of cylinders is an even number.

  According to the first aspect of the present invention, the rotation angle position of the input shaft beyond the top dead center from the compression stroke to the expansion stroke is used as the reference position, and the rotation angle difference between the reference positions of the cylinders adjacent in the combustion order is By configuring the cam disk to be different from a predetermined angle in the circumferential direction of the input shaft (cam disk arrangement phase angle), even if torque fluctuations due to the operation stroke of the internal combustion engine are applied to the input shaft, the torque The strength required for each lever crank mechanism can be ensured by designing the strength of the plurality of lever crank mechanisms corresponding to the fluctuations. Further, as compared with the case where the strength is uniformly set for a plurality of lever crank mechanisms, it is possible to realize a reduction in size and weight or a reduction in cost.

  According to the second aspect of the present invention, the input shaft with the rotation angle position at the reference position is subjected to a larger torque than the other rotation angle positions, so that the strength of the lever crank mechanism close to the reference position is increased. By doing so, the strength of the driving force from the internal combustion engine can be optimized.

  According to the invention of claim 3, by making the width of the connecting rod of the lever crank mechanism to which a large torque is applied wider than the width of the connecting rod of the other lever crank mechanism, the sectional area perpendicular to the torque transmission path is changed. Can cope with the magnitude of torque.

  According to the invention of claim 4, since the torque that the input shaft receives from the crankshaft of the internal combustion engine becomes the maximum after the start of the expansion stroke in the operation stroke of the internal combustion engine, the strength of the lever crank mechanism close to the reference position. It leads to setting to raise.

  According to the invention of claim 5, when the number of cylinders of the internal combustion engine is an odd number, the phase of the output torque to each lever crank mechanism is reversed, and the setting of the strength of each lever crank mechanism becomes complicated. The number of cylinders in the engine is desirably an even number.

It is a block diagram which shows the internal structure of the vehicle containing the continuously variable transmission of this embodiment. It is sectional drawing of the axial direction which shows embodiment of the continuously variable transmission of this invention. It is a schematic diagram which shows the turning radius adjustment mechanism, connecting rod, and rocking | fluctuation link of the continuously variable transmission shown in FIG. 2 from an axial direction. It is a schematic diagram which shows the change of the rotation radius of the rotation radius adjustment mechanism of the continuously variable transmission of FIG. FIG. 3 is a schematic diagram showing the relationship between the change in the rotation radius of the rotation radius adjustment mechanism of the continuously variable transmission of FIG. 2 and the swing angle of the swing motion of the swing link, where (a) shows the maximum rotation radius; ) Shows the case where the turning radius is medium, and (c) shows the case where the turning radius is small. It is a graph which shows the change of the angular velocity of the rocking | fluctuation link with respect to the change of the rotation radius of the rotation radius adjustment mechanism of the continuously variable transmission of FIG. 3 is a graph showing a state where an output shaft is rotated by a lever crank mechanism of the continuously variable transmission of FIG. 2. FIG. 3 is a schematic diagram showing a relationship between cam disks of a lever crank mechanism of the continuously variable transmission shown in FIG. 2, (a) is an arrangement relationship of each cam disk, (b) is a phase relationship of each cam disk, ( c) shows the inter-axis distance relationship of each cam disk. It is typical explanatory drawing which shows the relationship between the torque fluctuation of the output torque of a 4-cylinder 4 cycle engine, and the six lever crank mechanisms with which the continuously variable transmission was equipped. It is typical explanatory drawing which shows the relationship between the torque fluctuation of the output torque of a 2 cylinder 4 cycle engine, and the six lever crank mechanisms with which the continuously variable transmission was equipped. It is sectional drawing of the axial direction of the continuously variable transmission which concerns on a modification.

  Hereinafter, embodiments of the continuously variable transmission of the present invention will be described.

  FIG. 1 is a block diagram showing an internal configuration of a vehicle including a continuously variable transmission according to this embodiment. The continuously variable transmission 1 mounted on the vehicle shown in FIG. 1 transmits the driving force from the engine E to the drive wheels W, W via the left and right axles.

  The continuously variable transmission 1 according to the present embodiment is a four-bar linkage type continuously variable transmission, and the transmission ratio i (i = rotational speed of the input shaft / rotational speed of the output shaft) is set to infinity (∞). It is a kind of transmission that can make the rotation speed of the output shaft “0”, so-called IVT (Infinity Variable Transmission).

  First, with reference to FIG.2 and FIG.3, the structure of the continuously variable transmission 1 of this embodiment is demonstrated.

  The continuously variable transmission 1 according to the present embodiment includes an input shaft 2, an output shaft 3, and six turning radius adjustment mechanisms 4. The continuously variable transmission 1 is housed in a transmission case 21. The transmission case 21 has one end wall portion 21a, the other end wall portion 21b disposed opposite to the one end wall portion 21a and fixed to the engine ENG, the outer edge of the one end wall portion 21a, and the other end wall portion 21b. It is formed by the peripheral wall part 21c which connects an outer edge. The one end wall 21a and the other end wall 21b are formed with an input shaft side opening for supporting the input shaft 2 and an output shaft side opening for supporting the output shaft 3, respectively. An input shaft bearing 22 and an output shaft bearing 23 are fitted into the input shaft side opening and the output shaft side opening.

  The input shaft 2 is a hollow member, and rotates about the rotation center axis P <b> 1 of the input shaft 2 by receiving a rotational driving force from the engine E.

  The output shaft 3 is disposed in parallel with the input shaft 2 and transmits rotational power to a drive unit such as a drive wheel W of the vehicle via a differential gear D, an axle, or the like.

  Each of the turning radius adjusting mechanisms 4 is provided so as to rotate about the rotation center axis P1 of the input shaft 2, and includes a cam disk 5 as a cam part, a rotating disk 6 as a rotating part, and a pinion shaft 7. Have.

  The cam disks 5 have a disk shape, and are provided in pairs on the input shaft 2 so as to be eccentric from the rotation center axis P <b> 1 of the input shaft 2 and rotate integrally with the input shaft 2. Each set of cam disks 5 is arranged so as to make a round in the circumferential direction of the input shaft 2 with six sets of cam disks 5.

  The rotating disk 6 has a disk shape in which a receiving hole 6a is provided at a position eccentric from the center thereof, and is rotatably fitted to the cam disk 5 one by one through the receiving hole 6a. doing.

  The center of the receiving hole 6a of the rotating disk 6 is a distance Ra from the rotation center axis P1 of the input shaft 2 to the center P2 of the cam disk 5 (center of the receiving hole 6a) and the center P2 of the cam disk 5 to the rotating disk 6. The distance Rb to the center P3 is the same. The receiving hole 6 a of the rotating disk 6 is provided with an internal tooth 6 b at a position between the pair of cam disks 5.

  The pinion shaft 7 is disposed concentrically with the input shaft 2 in the hollow input shaft 2 and is rotatable relative to the input shaft 2. Further, external teeth 7 a are provided on the outer periphery of the pinion shaft 7. External teeth 7 a provided on the outer periphery of the pinion shaft 7 mesh with internal teeth 6 b provided on the inner periphery of the receiving hole 6 a of the rotary disk 6. Further, a differential mechanism 8 is connected to the pinion shaft 7.

  The differential mechanism 8 is configured as a planetary gear mechanism, and includes a sun gear 9, a first ring gear 10 connected to the input shaft 2, a second ring gear 11 connected to the pinion shaft 7, the sun gear 9 and the first ring gear 10. And a carrier 13 that pivotally supports a stepped pinion 12 including a small-diameter portion 12b meshing with the second ring gear 11 so as to rotate and revolve. The sun gear 9 of the differential mechanism 8 is connected to a rotating shaft 14a of an adjustment drive source 14 composed of an electric motor for the pinion shaft 7.

  Therefore, when the rotational speed of the adjusting drive source 14 is the same as the rotational speed of the input shaft 2, the sun gear 9 and the first ring gear 10 rotate at the same speed, so that the sun gear 9, the first ring gear 10, the second gear The four elements of the ring gear 11 and the carrier 13 are locked so as not to rotate relative to each other, and the pinion shaft 7 connected to the second ring gear 11 rotates at the same speed as the input shaft 2.

  When the rotational speed of the adjusting drive source 14 is made slower than the rotational speed of the input shaft 2, the rotational speed of the sun gear 9 is Ns, the rotational speed of the first ring gear 10 is NR1, and the gear ratio between the sun gear 9 and the first ring gear 10 ( When j is the number of teeth of the first ring gear 10 / the number of teeth of the sun gear 9, the rotation speed of the carrier 13 is (j · NR1 + Ns) / (j + 1). Further, the gear ratio between the sun gear 9 and the second ring gear 11 ((number of teeth of the second ring gear 11 / number of teeth of the sun gear 9) × (number of teeth of the large diameter portion 12a of the stepped pinion 12 / number of teeth of the small diameter portion 12b). ) Is k, the rotation speed of the second ring gear 11 is {j (k + 1) NR1 + (k−j) Ns} / {k (j + 1)}.

  Therefore, when the rotational speed of the adjusting drive source 14 is made slower than the rotational speed of the input shaft 2, the rotational speed of the input shaft 2 to which the cam disk 5 is fixed and the rotational speed of the pinion shaft 7 are the same. In some cases, the rotating disk 6 rotates together with the cam disk 5. On the other hand, when there is a difference between the rotation speed of the input shaft 2 and the rotation speed of the pinion shaft 7, the rotating disk 6 rotates around the center P <b> 2 of the cam disk 5.

  As shown in FIG. 3, the rotating disk 6 is eccentric with respect to the cam disk 5 so that the distance Ra from P1 to P2 and the distance Rb from P2 to P3 are the same. Therefore, the center P3 of the rotating disk 6 is positioned concentrically with the rotation center axis P1 of the input shaft 2, and the distance between the rotation center axis P1 of the input shaft 2 and the center P3 of the rotating disk 6, that is, the amount of eccentricity R1 is expressed as “ It can also be set to “0”.

  A connecting rod 15 is rotatably fitted around the periphery of the rotating radius adjusting mechanism 4, specifically, the rotating disk 6 of the rotating radius adjusting mechanism 4.

  The connecting rod 15 has a large-diameter large-diameter annular portion 15a at one end, and a small-diameter annular portion 15b having a smaller diameter than the diameter of the large-diameter annular portion 15a at the other end. The large-diameter annular portion 15a of the connecting rod 15 is externally fitted to the rotary disk 6 via a connecting rod bearing 16 formed of a ball bearing.

  A swing link 18 is pivotally supported on the output shaft 3 via a one-way clutch 17 as a one-way rotation prevention mechanism.

  The one-way clutch 17 fixes the swing link 18 with respect to the output shaft 3 when trying to rotate to one side around the rotation center axis P4 of the output shaft 3, and outputs when trying to rotate to the other side. The swing link 18 is idled with respect to the shaft 3.

  The swing link 18 is provided with a swing end portion 18a, and the swing end portion 18a is provided with a pair of projecting pieces 18b formed so as to sandwich the small-diameter annular portion 15b in the axial direction. Yes. The pair of projecting pieces 18b are formed with through holes 18c corresponding to the inner diameter of the small-diameter annular portion 15b. The connecting rod 15 and the swing link 18 are connected by inserting the connecting pin 19 into the through hole 18c and the small-diameter annular portion 15b. Further, the swing link 18 is provided with an annular portion 18d.

  In the present embodiment, the one-way clutch 17 is used as the one-way rotation prevention mechanism, but the one-way rotation prevention mechanism used in the continuously variable transmission of the present invention is not limited to this, and is output from the swing link 18. You may comprise with the two-way clutch (two-way clutch) comprised so that the rotation direction with respect to the output shaft 3 of the rocking | fluctuation link 18 which can transmit a torque to the axis | shaft 3 is changeable.

  Next, the lever crank mechanism of the continuously variable transmission 1 according to the present embodiment will be described with reference to FIGS.

  As shown in FIG. 3, in the continuously variable transmission 1 according to the present embodiment, the turning radius adjusting mechanism 4, the connecting rod 15, and the swing link 18 constitute a lever crank mechanism 20 (four-bar linkage mechanism). ing.

  The lever crank mechanism 20 converts the rotational motion of the input shaft 2 into the swing motion of the swing link 18. As shown in FIG. 2, the continuously variable transmission 1 of this embodiment includes a total of six lever crank mechanisms 20.

  In this lever crank mechanism 20, when the eccentric amount R1 of the turning radius adjusting mechanism 4 is not "0", when the input shaft 2 and the pinion shaft 7 are rotated at the same speed, each connecting rod 15 has a phase of 60 degrees. While changing, the swing link 18 is swung by alternately repeating pushing between the input shaft 2 and the output shaft 3 toward the output shaft 3 and pulling toward the input shaft 2.

  Since the one-way clutch 17 is provided between the swing link 18 and the output shaft 3, the swing link 18 is pushed or pulled and the swing link 18 is swung. The dynamic link 18 is fixed, and the force of the swinging motion of the swinging link 18 is transmitted to the output shaft 3 to rotate the output shaft 3. In the other case, the swinging link 18 is idled to the output shaft 3. The force of the swing motion of the swing link 18 is not transmitted. Since the six turning radius adjusting mechanisms 4 are arranged by changing the phase by 60 degrees, the output shaft 3 is sequentially rotated by the six turning radius adjusting mechanisms 4.

  Moreover, in the continuously variable transmission 1 of this embodiment, as shown in FIG. 4, the rotation radius of the rotation radius adjustment mechanism 4, that is, the eccentric amount R1 is adjustable.

  FIG. 4A shows a state in which the eccentric amount R1 is set to “maximum”, and the rotation center axis P1 of the input shaft 2, the center P2 of the cam disk 5, and the center P3 of the rotation disk 6 are aligned. The pinion shaft 7 and the rotating disk 6 are located. In this case, the gear ratio i is minimized. FIG. 4B shows a state where the eccentric amount R1 is set to “medium” which is smaller than that in FIG. 4A, and FIG. 4C shows that the eccentric amount R1 is smaller than that in FIG. Is shown. The gear ratio i is “medium” which is larger than the gear ratio i in FIG. 4A in FIG. 4B, and “large” which is larger than the gear ratio i in FIG. 4B in FIG. Become. FIG. 4D shows a state where the eccentricity R1 is “0”, and the rotation center axis P1 of the input shaft 2 and the center P3 of the rotating disk 6 are located concentrically. In this case, the gear ratio i is infinite (∞).

  FIG. 5 is a schematic diagram showing the relationship between the rotation radius of the rotation radius adjusting mechanism 4 of this embodiment, that is, the change in the eccentricity R1 and the swing angle (swing range) of the swinging motion of the swing link 18. FIG.

  FIG. 5A shows the case where the eccentric amount R1 is “maximum” in FIG. 4A (when the gear ratio i is the minimum), and FIG. 5B shows the amount of eccentricity R1 in FIG. FIG. 5C shows the case where the eccentric amount R1 is “small” in FIG. 4C (when the gear ratio i is large). The swing range θ2 of the swing link 18 with respect to the rotational movement of the turning radius adjusting mechanism 4 is shown. Here, the distance from the rotation center axis P4 of the output shaft 3 to the connecting point of the connecting rod 15 and the swinging end portion 18a, that is, the center P5 of the connecting pin 19, is the length R2 of the swinging link 18.

  As is apparent from FIG. 5, as the eccentric amount R1 becomes smaller, the swing range θ2 of the swing link 18 becomes narrower, and when the eccentric amount R1 becomes “0”, the swing link 18 swings. Stops moving.

  6 shows the angular velocity associated with the change in the eccentric amount R1 of the rotating radius adjusting mechanism 4 with the phase θ1 of the rotating radius adjusting mechanism 4 of the continuously variable transmission 1 as the horizontal axis and the angular velocity ω of the swing link 18 as the vertical axis. It is a figure which shows the relationship of the change of (omega).

  As can be seen from FIG. 6, the angular velocity ω of the swing link 18 increases as the eccentric amount R1 increases (the transmission ratio i decreases).

  Further, FIG. 7 shows each oscillation with respect to the phase θ1 of the rotation radius adjustment mechanism 4 when the six rotation radius adjustment mechanisms 4 are rotated (when the input shaft 2 and the pinion shaft 7 are rotated at the same speed). It is a figure which shows angular velocity (omega) of the link.

  It can be seen from FIG. 7 that the output shaft 3 is smoothly rotated by the six lever crank mechanisms 20.

  Further, as shown in FIG. 2, the continuously variable transmission 1 of the present embodiment includes six lever crank mechanisms 20.

  In the continuously variable transmission 1 of the present embodiment, as shown in FIG. 8 (a), the turning radius adjusting mechanism 4 of each of the six lever crank mechanisms 20 is provided on the side opposite to the engine. That is, in order from the adjustment drive source (ACT) 14 side, the first to sixth turning radius adjustment mechanisms 4a to 4f are used, and each set of cam disks 5 included in each of them is referred to as the first to sixth cam disks 5a. ~ 5f. Further, the six lever crank mechanisms 20 are first to sixth lever crank mechanisms 20a to 20f in order from the adjustment drive source 14 side.

  Then, as shown in FIG. 8B, the phases of the first cam disk 5a and the second cam disk 5b are shifted by 60 ° counterclockwise in FIG. 8B. Further, the phases of the second cam disk 5b and the third cam disk 5c are shifted by 60 ° counterclockwise in FIG. 8B. The phases of the third cam disk 5c and the fourth cam disk 5d are shifted by 60 ° counterclockwise in FIG. 8B. The phases of the fourth cam disk 5d and the fifth cam disk 5e are shifted by 60 ° counterclockwise in FIG. 8B. The phases of the fifth cam disk 5e and the sixth cam disk 5f are shifted by 60 ° counterclockwise in FIG. 8B. Note that the rotation direction of the crankshaft of the engine is counterclockwise in FIG.

  Further, as shown in FIG. 8C, the cam disks 5 a to 5 f are arranged at equal intervals in the axial direction of the input shaft 2.

  In the continuously variable transmission 1 according to the present embodiment, the phase of the first to sixth cam disks 5a to 5f, that is, the phase and arrangement of the first to sixth turning radius adjusting mechanisms 4a to 4f are set in such a relationship. Accordingly, the centrifugal forces generated in the respective turning radius adjusting mechanisms 4 cancel each other, and vibrations generated when the respective turning radius adjusting mechanisms 4 rotate are suppressed.

  Next, a configuration in which the optimum fatigue strength design is applied to the continuously variable transmission 1 when the engine E is a four-cylinder four-cycle engine will be described with reference to FIG. FIG. 9 is a schematic explanatory view showing the relationship between the torque fluctuation of the output torque of the four-cylinder four-cycle engine and the six lever crank mechanisms 20 provided in the continuously variable transmission 1.

  The horizontal axis of the graph shown in FIG. 9 indicates the crank angle (crankshaft rotation angle) of the four-cylinder four-cycle engine, and the vertical axis indicates the output torque of the engine. The output torque of the four-cylinder four-cycle engine indicates the torque transmitted from the piston to the input shaft 2 via the crankshaft during each operation stroke in the four cylinders C1 to C4). Further, a four-cylinder four-cycle engine (hereinafter simply referred to as “engine”) completes a series of operation strokes with a crank angle of 720 degrees, but in the example shown in FIG. 9, each cylinder C1 to C4 has 180 degrees. The expansion strokes are set so as to reach an equal phase interval. The cylinders C1 to C4 are set so as to reach the expansion stroke in the order of the cylinders C1, C2, C3, and C4, which are not related to the physical arrangement, according to the ignition order as a countermeasure against engine vibration.

  In the graph shown in FIG. 9, the crank angle (rotational angle position of the input shaft 2) when the piston of the cylinder C1 is located at the top dead center (TDC) is 0 degree, and the cylinder C1 expands next to the cylinder C1. The crank angle when the piston of the cylinder C2 that reaches the stroke is located at the top dead center is 180 degrees, and the crank angle when the piston of the cylinder C3 that reaches the expansion stroke next to the cylinder C2 is located at the top dead center is 360 degrees. The crank angle when the piston of the cylinder C4 that reaches the expansion stroke next to the cylinder C3 is located at the top dead center is 540 degrees.

  Further, the crank angle t1 at which the output torque obtained from the cylinder C1 is maximized at the time when the top dead center (crank angle = 0 degree) from the compression stroke to the expansion stroke in the cylinder C1 is exceeded is defined as the reference of the cylinder C1. Position. Similarly, the crank angle t2 at which the output torque obtained from the cylinder C2 is maximized when the top dead center (crank angle = 180 degrees) from the compression stroke to the expansion stroke in the cylinder C2 is exceeded is determined as The reference position. Further, the crank angle t3 at which the output torque obtained from the cylinder C3 becomes the maximum at the time when the top dead center (crank angle = 360 degrees) from the compression stroke to the expansion stroke in the cylinder C3 is exceeded is determined as the reference of the cylinder C3. Position. Further, the crank angle t4 at which the output torque obtained from the cylinder C4 becomes the maximum at the time when the top dead center (crank angle = 540 degrees) from the compression stroke to the expansion stroke in the cylinder C4 is exceeded is defined as the reference of the cylinder C4. Position.

  As shown in FIG. 8, the rotational radius adjusting mechanisms 4a to 4f in the six lever crank mechanisms 20 provided in the continuously variable transmission 1 described above are arranged on the input shaft 2 with a phase changed every 60 degrees. Yes. Therefore, in the present embodiment, the output torque from the cylinder C1 is mainly transmitted by the first lever crank mechanism 20a to the third lever crank mechanism 20c, and the output torque from the cylinder C2 is mainly transmitted to the fourth lever crank mechanism 20d to 20d. The output torque from the cylinder C3 is transmitted mainly by the first lever crank mechanism 20a to the third lever crank mechanism 20c, and the output torque from the cylinder C4 is mainly transmitted by the fifth lever crank mechanism 20f. 20d to the sixth lever crank mechanism 20f.

  However, as shown in FIG. 9, the relative angle of the eccentric angle of the cam disk 5 with respect to the reference position (crank angle t1) of the cylinder C1 is the smallest in the first lever crank mechanism 20a, and the second lever crank mechanism 20b is the next. The small third lever crank mechanism 20c is the largest. The relative angle of the eccentric angle of the cam disk 5 with respect to the reference position (crank angle t3) of the cylinder C3 is the smallest for the first lever crank mechanism 20a, the second for the second lever crank mechanism 20b, and the third lever crank mechanism. 20c is the largest. For this reason, the magnitude of the output torque received from the cylinders C1 and C3 decreases in the order of the first lever crank mechanism 20a, the second lever crank mechanism 20b, and the third lever crank mechanism 20c. Thus, since the magnitude | size of the load which the six lever crank mechanisms 20 receive differs, in this embodiment, the width | variety of the connecting rod 15 of the lever crank mechanism 20 is comprised widely according to the magnitude | size of load. Note that if the connecting rod 15 is wide, the cross-sectional area perpendicular to the torque transmission path increases, and the strength against the load increases.

  Thus, the widths of the connecting rods 15 of the first to third lever crank mechanisms 20a to 20c are different from each other, and the first and second lever crank mechanisms 20a are larger than the width of the connecting rod 15 of the third lever crank mechanism 20c. , 20b, and the connecting rod 15 of the second lever crank mechanism 20b is wider than the connecting rod 15 of the second lever crank mechanism 20b.

  The relative angle of the eccentric angle of the cam disk 5 with respect to the reference positions (crank angles t2, t4) of the cylinders C2, C4 is the smallest for the fourth lever crank mechanism 20d, the second for the fifth lever crank mechanism 20e, and the sixth lever. The crank mechanism 20f is the largest. Therefore, the widths of the connecting rods 15 of the fourth to sixth lever crank mechanisms 20d to 20f are the same as the above relationship. That is, the widths of the connecting rods 15 of the fourth to sixth lever crank mechanisms 20d to 20f are different from each other according to the magnitude of the load, and are fifth and fifth than the width of the connecting rod 15 of the fourth lever crank mechanism 20d. The width of each connecting rod 15 of the six lever crank mechanisms 20e, 20f is made wider, and the width of the connecting rod 15 of the fourth lever crank mechanism 20d is wider than the width of the connecting rod 15 of the fifth lever crank mechanism 20e. doing.

  As described above, in the present embodiment, the width of the connecting rod 15 varies depending on the magnitude of the output torque from the engine received by the six lever crank mechanisms 20 provided in the continuously variable transmission 1. That is, the width of the connecting rod 15 of the lever crank mechanism 20 that receives a large load is wide, and the width of the connecting rod 15 of the lever crank mechanism 20 that receives a small load is narrow. Thus, the strength required for each lever crank mechanism 20 can be ensured because the widths of the connecting rods 15 are different. Further, since the lever crank mechanism 20 including the narrow connecting rod 15 is thin and light, compared to the case where the strength is uniformly set for all the lever crank mechanisms 20, a reduction in size and weight or a reduction in cost can be realized.

  The continuously variable transmission 1 of the present embodiment can be configured to have an optimum fatigue strength design for a two-cylinder four-cycle engine. FIG. 10 is a schematic explanatory view showing the relationship between the torque fluctuation of the output torque of the two-cylinder four-cycle engine and the six lever crank mechanisms 20 provided in the continuously variable transmission 1. As shown in FIG. 10, the output torque from the cylinder C1 is transmitted by the first leverage crank mechanism 20a to the sixth leverage crank mechanism 20f, and the output torque from the cylinder C2 is also transmitted from the first leverage crank mechanism 20a to the sixth leverage crank mechanism. 20f is transmitted.

  In this case, as shown in FIG. 10, the relative angle of the eccentric angle of the cam disk 5 with respect to the reference position (eg, crank angle t1) of one cylinder (eg, cylinder C1) is the smallest in the first lever crank mechanism 20a. The second lever crank mechanism 20b is the next smallest, and the third to sixth lever crank mechanisms 20c to 20f are large. For this reason, the magnitude of the output torque received from the one cylinder decreases in the order of the first lever crank mechanism 20a, the second lever crank mechanism 20b, and the third to sixth lever crank mechanisms 20c to 20f. However, the relative angle of the eccentric angle of the cam disk 5 in the sixth lever crank mechanism 20f with respect to the reference position of the one cylinder is the largest, but the reference position (for example, the crank angle t2) of the other cylinder (for example, the cylinder C2). The relative angle to is small. Accordingly, the sixth lever crank mechanism 20f generates an output torque that is relatively larger than the output torque from one cylinder where the first lever crank mechanism 20a and the second lever crank mechanism 20b receive a relatively large output torque. Receive from.

  As described above, even when a two-cylinder four-cycle engine is used, the loads received by the six lever crank mechanisms 20 are different from each other. Therefore, the width of the connecting rod 15 of the lever crank mechanism 20 depends on the load. Composed.

  In the above description, the width of the connecting rod 15 varies depending on the magnitude of the output torque from the engine received by the lever crank mechanism 20, but the strength of the lever crank mechanism 20 with respect to the load is changed using materials having different rigidity. May be. Further, not only the width of the connecting rod 15 is different, but also the one-way clutch 17 provided between the swing link 18 and the output shaft 3 shown in FIG. Further, if the number of cylinders of the engine E is an odd number, the phase of the output torque to each lever crank mechanism is reversed and the setting of the strength of each lever crank mechanism becomes complicated, so the number of cylinders of the engine E is an even number. It is desirable. The number of lever crank mechanisms 20 provided in the continuously variable transmission 1 is not limited to six, and may be two or more. The engine E is not limited to 4 cycles, and may be 2 cycles or 6 cycles. However, in relation to the number of cylinders of the engine E, the crank angle required for a series of operation strokes in each cylinder, and the number of lever crank mechanisms 20 provided in the continuously variable transmission 1, a plurality of lever crank mechanisms 20 are connected to the engine E from the engine E. Excludes configurations where output torque is equally applied. That is, it is only necessary that the rotational angle difference between the reference positions of the plurality of cylinders is different from the predetermined angle difference provided in the circumferential direction of the input shaft 2 of the cam disk 5 in the lever crank mechanism 20.

  The embodiments of the present invention have been described above, but various design changes can be made without departing from the scope of the present invention.

  For example, as shown in FIG. 11, the continuously variable transmission 1 is provided with a cylindrical surface 7e centered on the rotation center axis P1 on the cam disk 5 positioned substantially at the center of the input shaft 2, and the cylindrical surface 7e is a peripheral wall. You may comprise so that it may rotatably support via the intermediate bearing 26 in the partition part 21d extended from the part 21c. In this case, an intermediate bearing 27 that supports the output shaft 3 is provided on the output shaft 3 side of the partition wall 21d, and the link mechanism 40, the first relative angle sensor 48, and the second relative angle sensor 49 are provided as output shaft bearings. 23 or at least one of the swing links 18 adjacent to the intermediate bearing 27.

DESCRIPTION OF SYMBOLS 1 Continuously variable transmission 2 Input shaft 3 Output shaft 4 Turning radius adjustment mechanism 5 Cam disk 6 Rotating disk 6a Receiving hole 6b Inner tooth 7 Pinion shaft 7a Outer tooth 8 Differential mechanism 9 Sun gear 10 First ring gear 11 Second ring gear 12 Stage Pinion 12a Large-diameter portion 12b Small-diameter portion 13 Carrier 14 Adjustment drive source 14a Rotating shaft 15 Connecting rod 15a Large-diameter annular portion 15b Small-diameter annular portion 16 Connecting rod bearing 17 One-way clutch 18 Oscillating link 18a Oscillating end 18b Projection piece 18c Through hole 18d Annular portion 19 Connecting pin 20 Lever crank mechanism 21 Transmission case 21a One end wall portion 21b Other end wall portion 21c Peripheral wall portion 22 Input shaft bearing 23 Output shaft bearing 26 Intermediate bearing 27 Intermediate bearings C1, C2, C3, C4 Cylinder E Engine P1 Rotation center axis P2 of input shaft 2 Center of cam disk 5 3 Center P4 of the rotating disk 6 Rotation center axis P5 of the output shaft 3 Center Ra of the connecting pin 19 Distance Rb between the P1 and P2 Distance P1 between the P2 and P3 Distance R1 between the P1 and P3 )
R2 Distance between P4 and P5 (length of swing link 18)
θ1 Phase of turning radius adjusting mechanism 4 θ2 Swing range of rocking link 18

Claims (5)

An input shaft to which driving force from an internal combustion engine having a plurality of cylinders mounted on a vehicle is transmitted;
An output shaft disposed parallel to the input shaft;
An eccentric mechanism having a cam disk eccentrically provided on the input shaft, a rotating disk provided eccentrically with respect to the cam disk, and a swing link pivotally supported by the output shaft; A plurality of lever crank mechanisms for converting the rotational motion of the input shaft into the swing motion of the swing link;
The swing link is fixed to the output shaft when the swing link is about to rotate to one side around the output shaft, and the swing shaft is fixed to the output shaft when the swing link is about to rotate to the other side. A continuously variable transmission comprising: a one-way rotation prevention mechanism that idles the swing link;
The cam disks in the eccentric mechanisms of the plurality of lever crank mechanisms are eccentrically arranged at predetermined angles in the circumferential direction of the input shaft,
When the rotation angle position of the input shaft at the time of exceeding the top dead center in the stroke from the compression stroke to the expansion stroke in each cylinder of the internal combustion engine is used as the reference position, the reference of each cylinder adjacent in the combustion order The rotational angle difference between the positions is configured to be different from the predetermined angle,
At least one of the plurality of lever crank mechanisms is a continuously variable transmission having a strength against a driving force from the internal combustion engine different from that of other lever crank mechanisms. The continuously variable transmission according to claim 1,
Among the plurality of lever crank mechanisms, the lever crank mechanism having the smallest relative angle between the rotation angle of the input shaft at the reference position and the eccentric angle of the cam disk with respect to the input shaft is driven from the internal combustion engine. A continuously variable transmission that is stronger than other lever crank mechanisms. The continuously variable transmission according to claim 1 or 2,
The lever crank mechanism has a large-diameter annular portion that is rotatably fitted to the eccentric mechanism at one end portion, and a connecting rod that is connected to the swing end portion of the swing link at the other end portion. Have
A continuously variable transmission in which a width of the connecting rod of the lever crank mechanism having a high strength is wider than a width of the connecting rod of the lever crank mechanism having a low strength. The continuously variable transmission according to any one of claims 1 to 3,
The reference position is the rotational angle position of the input shaft after the expansion stroke starts, and the torque received by the input shaft at the reference position is the maximum torque obtained in the stroke of each cylinder of the internal combustion engine. There is a continuously variable transmission. A continuously variable transmission according to any one of claims 1 to 4,
The continuously variable transmission, wherein the number of the plurality of cylinders is an even number.

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