JP2006125257A - Parallel multi-cylinder four-cycle engine and saddle-riding type vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To achieve extremely favorable operating feeling and compactization. <P>SOLUTION: The parallel multi-cylinder four-cycle engine is equipped with a crank shaft 24 having a crank pin for supporting a connecting rod provided to each cylinder and a crankweb for supporting the crank pin, and a balancer shaft 32 arranged parallel to the crank shaft 24 and provided with a balance weight to eliminate couple of forces generated by primary inertia. The length of the balancer shaft 32 is made shorter than the space between crankwebs of both ends on the crank shaft 24 so as to arrange the balancer shaft 32 deflected to one side in a direction of the crank shaft. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a parallel multi-cylinder four-cycle engine having a balancer shaft that eliminates a couple due to primary inertia and a straddle-type vehicle equipped with a parallel multi-cylinder four-cycle engine.

Conventionally, in motorcycles, etc., a two-plane crank type engine has been proposed in which the combustion pressure of the engine is felt by the body in terms of vibration amplitude, S / N ratio and frequency, acceleration feeling is improved, and driving feeling is good. (For example, Patent Documents 1 and 2).
Japanese Examined Patent Publication No. 7-76577 (pages 1 to 15 and FIGS. 1 to 13) Japanese Patent Laid-Open No. 9-250597 (pages 1 to 6, FIGS. 1 to 6)

  In Patent Document 1, the crankshaft and the main shaft are arranged in parallel, the rotation of the crankshaft is transmitted to the main shaft through gears provided on both shafts, and the balancer shaft is also connected to the crankshaft through another gear. Since the rotation is transmitted, a plurality of rotation transmission parts are formed on the crankshaft, or a balancer shaft is provided at a distance from the main shaft, making it difficult to make the entire engine compact. It was. Further, Patent Document 2 has a problem that the balancer shaft is provided on the side of the crankshaft that is the widest of the engines, and the entire engine cannot be made compact.

  The present invention has been made in view of the above points, and an object of the present invention is to provide a compact parallel multi-cylinder four-cycle engine and a straddle-type vehicle having a very good driving feeling.

  In order to solve the above-described problems and achieve the object, the present invention is configured as follows.

The invention according to claim 1 is a crankshaft having a crankpin that supports a connecting rod provided for each cylinder, and a crank web that supports the crankpin;
A parallel multi-cylinder four-cycle engine that is arranged in parallel with the crankshaft and has a balance weight and eliminates the couple due to primary inertia;
The length of the balancer shaft is smaller than the distance between the crank webs at both ends on the crank shaft;
The parallel multi-cylinder four-cycle engine is characterized in that the balancer shaft is displaced to one side in the crankshaft direction.

The invention according to claim 2 is a crankshaft having a crankpin that supports a connecting rod provided for each cylinder, and a crank web that supports the crankpin;
A parallel multi-cylinder four-cycle engine that is arranged in parallel with the crankshaft and has a balance weight and eliminates the couple due to primary inertia;
The balancer shaft has a plurality of balance weights,
The balance weight at one end is provided at a position corresponding to the crank web of the cylinder at one end,
In the parallel multi-cylinder four-cycle engine, the balance weight at the other end is arranged at a position corresponding to a crank web of a cylinder adjacent to the cylinder at the other end.

  The invention according to claim 3 is characterized in that the balancer shaft is supported so that a rotation locus of the balance weight provided on the balancer shaft and a rotation locus of the crank web overlap each other when viewed in the crankshaft direction. The parallel multi-cylinder four-cycle engine according to claim 1 or 2.

  According to a fourth aspect of the present invention, a power input portion for inputting the power extracted from the crankshaft is provided at an intermediate portion between the balance weights of the balancer shaft. The parallel multi-cylinder four-cycle engine described in 1.

  The invention according to claim 5 is characterized in that a power transmission portion for transmitting the power from the balancer shaft to the outside is provided at the end of the balancer shaft opposite to the side where the balancer shaft is displaced. The parallel multi-cylinder four-cycle engine according to any one of claims 3 to 3.

  The invention according to claim 6 is characterized in that the bearing of the balancer shaft is arranged corresponding to the bearing of the crankshaft. It is a 4-cycle engine.

  The invention according to claim 7 is the parallel multi-cylinder four-cycle engine according to claim 6, wherein the bearing of the balancer shaft is disposed between the balance weight and the power take-out portion.

  The invention according to claim 8 is the parallel multi-cylinder four-cycle engine according to claim 6, wherein bearings of the balancer shaft are arranged on both sides of the power take-out portion.

  A ninth aspect of the invention is a straddle-type vehicle equipped with the parallel multi-cylinder engine according to any one of the first to ninth aspects.

  With the above configuration, the present invention has the following effects.

  According to the first aspect of the present invention, the length of the balancer shaft is made smaller than the interval between the crank webs at both ends on the crankshaft, and is deviated to one side in the crankshaft direction so as to be adjacent to the crankshaft. A space can be secured next to the balancer shaft. By arranging a power transmission unit, a clutch, and the like in this space, the entire engine can be made compact while the driving feeling is very good.

  According to the second aspect of the invention, the balance weight at one end of the balancer shaft is provided at a position corresponding to the crank web of the cylinder at one end, and the balance weight at the other end is connected to the crank of the cylinder adjacent to the cylinder at the other end. By arranging it at a position corresponding to the web, a space can be secured next to the balancer shaft corresponding to the cylinder at the other end. By arranging a power transmission unit, a clutch, and the like in this space, the entire engine can be made compact while the driving feeling is very good.

  According to the third aspect of the present invention, in addition to the first or second aspect, the balancer shaft has a balance weight rotation locus provided on the balancer shaft and a crank web rotation locus overlapping in the crankshaft direction view. Thus, the distance between the balancer shaft and the crankshaft can be shortened, and the entire engine can be made compact.

  According to the fourth aspect of the present invention, in addition to the second or third aspect, since the power input portion for inputting the power extracted from the crankshaft is provided at the intermediate portion between the balance weights of the balancer shaft, Power can be extracted and input without lengthening the shaft or crankshaft.

  According to the fifth aspect of the present invention, in addition to any one of the first to third aspects, the power from the balancer shaft is transmitted to the end of the balancer shaft on the opposite side to the deflected side. Since the power transmission unit is provided, the entire engine can be made compact by arranging the power transmission unit in the space.

  According to the sixth aspect of the invention, in addition to any one of the first to fifth aspects, the bearing of the balancer shaft is arranged corresponding to the bearing of the crankshaft, so that the bearing support portion can be shared.

  According to the seventh aspect of the present invention, in addition to the sixth aspect, the balancer shaft bearing can be disposed between the balance weight and the power take-out portion to reliably support the balancer shaft.

  According to the eighth aspect of the present invention, in addition to the sixth aspect, the balancer shaft bearings can be disposed on both sides of the power take-out portion to reliably support the balancer shaft.

  According to the ninth aspect of the invention, the parallel multi-cylinder engine according to any one of the first to ninth aspects is mounted, and the entire engine can be made compact while the driving feeling is very good. It is a straddle-type vehicle.

  Hereinafter, embodiments of a parallel multi-cylinder four-cycle engine and a saddle-ride type vehicle according to the present invention will be described. The embodiment of the present invention shows the most preferable mode of the present invention, and the present invention is not limited to this. In this embodiment, a motorcycle is shown as a saddle-ride type vehicle, but the present invention can be similarly applied to an automatic tricycle or a buggy. Further, as a parallel multi-cylinder four-cycle engine, a four-cylinder four-cycle engine is shown, but the present invention can be similarly applied to a four-cycle engine such as a two-cylinder, three-cylinder, five-cylinder, or six-cylinder.

  1 is a side view of a motorcycle, FIG. 2 is a cross-sectional view of a parallel multi-cylinder four-cycle engine, FIG. 3 is a cross-sectional view showing the arrangement of a crankshaft, a balancer shaft, and a main shaft, and FIG. 4 is an arrangement of a crankshaft and a balancer shaft FIG. 5 is a side view showing the valve drive force transmission mechanism, FIG. 6 is a side view showing the cam shaft portion of the valve drive mechanism, and FIG. 7 is a side view showing the arrangement of the crankshaft and the balancer shaft. 8 is a view of the camshaft of the valve mechanism as viewed from the camshaft direction.

  In the motorcycle according to this embodiment, as shown in FIG. 1, an engine 1 is mounted on a body frame 2 of the motorcycle. A front fork 10 is turnably provided on the front side of the body frame 2, and a front wheel 11 is supported on the front fork 10. A rear arm 12 is pivotally supported on the rear side of the body frame 2 by a link mechanism 4 and a rear cushion 5 so as to be swingable up and down, and a rear wheel 13 is supported on the rear arm 12.

  The engine 1 is a four-cycle engine including four cylinders 3a to 3d, and these cylinders 3a to 3d are arranged in series so as to be inclined forward and orthogonal to the traveling direction of the vehicle. . An exhaust pipe 6 is connected to the front side of each of the cylinders 3 a to 3 d, and the exhaust pipe 6 extends rearward of the vehicle body through the front side and the lower side of the engine 1. An intake pipe 7 is connected to the rear side of the cylinders 3a to 3d, and an air cleaner 8 is connected to the intake pipe 7. Each of the four cylinders 3 a to 3 d of the engine 1 is provided with a fuel injection device 9.

  Next, as shown in FIGS. 2 to 8, the engine 1 has a structure in which a cylinder body 20 and a cylinder head 21 are stacked on a crankcase 19 that is divided into upper and lower parts, and a head cover 50 is covered on the upper surface of the cylinder head 21. belongs to. An oil pan 52 is provided below the crankcase 19.

  As shown in FIGS. 2 and 3, each of the cylinders 3 a to 3 d includes a cylinder body 20 and a cylinder head 21 mounted on a crankcase 19 including an upper case 17 and a lower case 18. Pistons 22a to 22d are slidably fitted in the cylinder holes 3a1 to 3d1 of the cylinders 3a to 3d. Combustion chambers 80 a to 80 d are formed between the heads of the pistons 22 a to 22 d and the recesses 21 a to 21 d of the cylinder head 21.

  In each of the combustion chambers 80a to 80d, three intake valve openings 81b, 81a and 81b and two exhaust valve openings 82a and 82a are opened. The left, right, and center intake valve openings 81b, 81b, and 81a are led out to the rear wall of the cylinder head 21 by a trifurcated intake passage 81c, and the left, right exhaust valve openings 82a and 82a are bifurcated exhaust passages. 82b leads to the front wall of the cylinder head 21.

  The left and right intake valve openings 81b and the central intake valve opening 81a are provided with left and right intake valves 83 and a central intake valve 84, respectively, and the exhaust valve opening 82a is provided with an exhaust valve 85. The valves 83 to 85 extend upward so as to form a predetermined clamping angle, and are provided by valve springs 86c1 and 86c2 interposed between retainers 86a1 and 86a2 fixed to the upper ends of the valves 86b1 and 86b2. Energized in the closing direction. An intake lifter 87a1 is mounted on the upper part of each intake valve 83, 84, 83, and this intake lifter 87a1 is in contact with the cams 88a1, 88a1, 88a1 of the intake side camshaft 88a. An exhaust lifter 87b1 is mounted on the upper portion of each exhaust valve 85, 85, and this exhaust lifter 87b1 is in contact with the cams 88b1, 88b1 of the exhaust side camshaft 88b.

  In this embodiment, the engine 1 is an OHC engine, and the valve operating mechanism G has an intake side cam shaft 88a and an exhaust side cam shaft 88b. By the rotation of the intake side cam shaft 88a and the exhaust side cam shaft 88b, In this configuration, intake and exhaust are performed. The intake side camshaft 88a and the exhaust side camshaft 88b are disposed in parallel with the crankshaft 24 and are rotatably supported by the bearing portion of the cylinder head 21. In the valve mechanism G of this embodiment, the intake side camshaft 88a is disposed above the intake valves 83 and 84, and the cams 88a1, 88a1 and 88a1 are moved to the intake valves 83 and 84 by the rotation of the intake side camshaft 88a. , 83 are driven via the intake lifter 87a1. Similarly, the exhaust side camshaft 88b is disposed close to the exhaust valves 85 and 85, and the cams 88b1 and 88b1 drive the exhaust valves 85 and 85 via the exhaust lifter 87b1 by the rotation of the exhaust side camshaft 88b. The valve mechanism G may drive the intake valves 83, 84, 83 and the exhaust valves 85, 85 by a rocker arm or the like.

  Pistons 22a to 22d are slidably fitted in the cylinder holes 3a1 to 3d1 of the cylinders 3a to 3d. Each piston 22a-22d is connected to crankpins 25a-25d of the crankshaft 24 via connecting rods 23a-23d. The crank pins 25a to 25d are installed between a pair of crank webs 26a to 26d of the crankshaft 24 corresponding to the cylinders 3a to 3d. The crankshaft 24 is rotatably supported by five bearing support portions 19a1 to 19e1 of the crankcase 19 via bearings 27a to 27e, respectively.

  A balancer shaft 32 is disposed in the crankcase 19 in parallel with the crankshaft 24. The balancer shaft 32 is rotatably supported by four bearing support portions 19a to 19d of the crankcase 19 via respective bearings 28a to 28d. The bearings 28 a to 28 d of the balancer shaft 32 are arranged corresponding to the bearings 27 a to 27 d of the crankshaft 24, and can share the four bearing support portions 19 a to 19 d of the crankcase 19.

  A driven gear 33 provided at the center portion of the balancer shaft 32 meshes with a drive gear 34 provided at the center portion of the crankshaft 24. The driven gear 33 and the drive gear 34 have the same diameter, and the crankshaft 24 and the balancer shaft 32 rotate in opposite directions at the same rotational speed. Balance weights 35 and 36 are provided on both sides of the driven gear 33 at portions from both ends of the balancer shaft 32.

  Further, in this embodiment, as shown in FIGS. 3 and 4, a power input portion C is provided in the middle portion between the balance weights 35 and 36 of the balancer shaft 32 to extract and input power from the crankshaft 24. The power input portion C is composed of a driven gear 33 that meshes with the drive gear 34. The drive gear 34 is formed integrally with the crank web 26b, and the driven gear 33 is provided by being press-fitted into the balancer shaft 32, so that power can be taken out without lengthening the balancer shaft 32 and the crankshaft 24. The bearing 28b of the balancer shaft 32 is disposed between the balance weight 35 and the power input portion C, and the bearing 28c is disposed between the balance weight 36 and the power input portion C, so that the balancer shaft 32 can be reliably supported. . Further, the bearings 28b and 28c of the balancer shaft 32 are arranged on both sides of the power input portion C, so that the balancer shaft 32 can be reliably supported.

  As shown in FIGS. 2 and 8, the balancer shaft 32 has one balance weight 35 of the balancer shaft 32 disposed between the crank webs 26a and 26a on both sides of the crank pin 25a of the cylinder 3a. The other balance weight 36 of the balancer shaft 32 is arranged corresponding to the crank webs 26c, 26c on both sides of the crank pin 25c of the cylinder 3d, and the radial front ends of the balance weights 35, 36 are the crank webs 26a, 26a and 26c. The crank webs 26c and 26c are configured to pass through the inner portion of the outer periphery of the crank webs 26c and 26c and rotate. As described above, the balancer shaft 32 is configured such that the rotation locus LO1 of the balance weights 35 and 36 provided on the balancer shaft 32 and the rotation locus LO2 of the crank webs 26a and 26a and the crank webs 26c and 26c overlap in the crankshaft direction view. Thus, the distance between the balancer shaft 32 and the crankshaft 24 can be shortened, and the entire engine can be made compact.

  In this embodiment, as shown in FIGS. 2 and 3, a valve driving force transmission mechanism H that transmits power from the balancer shaft 32 to the valve operating mechanism G is provided. In conjunction with the shaft 32, the intake side cam shaft 88a and the exhaust side cam shaft 88b of the valve mechanism G are rotated to perform intake and exhaust at a predetermined timing.

  The power input side H1 of the valve drive force transmission mechanism H is constituted by a cam drive gear 70 disposed between the balance weights 35 and 36 on the balancer shaft 32. The cam drive gear 70 is smaller in diameter than the driven gear 33 and has a speed reduction mechanism J that decelerates on the balancer shaft 32. The speed reduction mechanism J includes a cam drive gear 70 having a smaller diameter than the driven gear 33. The power output side H2 of the valve drive force transmission mechanism H includes an intake side driven gear 71 and an exhaust gas disposed on a cam shaft constituted by the intake side cam shaft 88a and the exhaust side cam shaft 88b of the valve mechanism G. A side driven gear 72 is used. The intake side driven gear 71 and the exhaust side driven gear 72 mesh with each other and rotate in conjunction with each other.

  This valve driving force transmission mechanism H has an intermediate gear group H3 between the power input side H1 and the power output side H2. The intermediate gear group H3 includes a first gear 73, a second gear 74, a third gear 75, and a fourth gear 76, which rotate in mesh with each other. The first gear 73 is formed on the rotary shaft 90, and the rotary shaft 90 is pivotally supported by a bearing 91a. The second gear 74 and the third gear 75 are formed on a rotating shaft 92, and the rotating shaft 92 is pivotally supported by a bearing 93a. The fourth gear 76 is rotatably supported by the gear case 94. The first gear 73 meshes with the cam drive gear 70, the fourth gear 76 meshes with the intake side driven gear 71, and the power of the balancer shaft 32 is transmitted from the cam drive gear 70 to the first gear 73, the second gear 74, and the third gear. It is transmitted to the gear 75 and the fourth gear 76 and further transmitted to the intake side driven gear 71 and the exhaust side driven gear 72 of the valve mechanism G to rotate the intake side cam shaft 88a and the exhaust side cam shaft 88b. In this embodiment, the valve driving force transmission mechanism H is configured by a gear group, but may be configured by a chain, a belt, or the like.

  As described above, the valve drive force transmission mechanism H that transmits power from the balancer shaft 32 to the valve mechanism G is provided, and power is not taken out from the crankshaft 24 to the valve mechanism G. An engine having a very good driving feeling can be made compact while ensuring driving of the driving force transmission mechanism H.

  Further, since the power input side H1 of the valve drive force transmission mechanism H is disposed between the balance weights 35 and 36 on the balancer shaft 32, the kinks of the balancer shaft 32 can be reduced, rotation fluctuation is small, and more stable. The driven valve drive force transmission mechanism H can be secured.

  Further, a power input portion C for inputting power from the crankshaft 24 is provided on the balancer shaft 32, and the cam drive gear 70 on the power input side H1 of the valve driving force transmission mechanism H is provided in the vicinity of the power input portion C. With this arrangement, it is possible to reduce the kinking of the balancer shaft 32, to reduce the rotational fluctuation, and to ensure more stable driving of the valve drive force transmission mechanism H. Further, since the power input portion C is disposed between the plurality of balance weights 35 and 36, the kinking of the balancer shaft 32 can be reduced, rotation fluctuation is small, and the valve driving force transmission mechanism H can be driven more stably. Can be secured. Further, the power input unit C can input the power extracted from the center of the crankshaft 24, thereby reducing the kinking of the crankshaft 24 and reducing the rotational fluctuation.

  Further, as shown in FIG. 2, the valve driving force transmission mechanism H is arranged outside the cylinder holes 3a1 to 3d1 as viewed in the cylinder axial direction, so that the interval between the cylinder holes 3a1 to 3d1 as shown in FIG. W1 can be packed and the engine can be made compact by shortening in the direction of the crankshaft.

  Further, as shown in FIG. 6, the power output side H2 of the valve drive force transmission mechanism H has a flange portion 88a2 formed on the intake side cam shaft 88a, and the intake side driven gear 71 is connected to the flange portion 88a2 by a bolt 88a3. Tightened and fixed. A flange portion 88b2 is formed on the exhaust-side cam shaft 88b, and the exhaust-side driven gear 72 is fastened and fixed to the flange portion 88b2 by a bolt 88b3. Thus, by arranging the intake side driven gear 71 and the exhaust side driven gear 72 on the cam shaft of the valve mechanism G without using any special member, the engine can be made compact by shortening in the cam shaft direction.

  In this embodiment, balance weights 35 and 36 are provided on the balancer shaft 32 asymmetrically with respect to the center of the balancer shaft 32. In the case of a balancer that eliminates the couple due to primary inertia, it can be arranged asymmetrically with respect to the center of the crank. Therefore, even if the position of the valve drive force transmission mechanism H is freely arranged, the balancer function is not impaired. Absent.

  The crank webs of the cylinders 3a to 3d each have a pair of crank webs 26a to 26d. The ratio of the reciprocating mass and the rotational mass of the pair of crank webs 26a of the first cylinder 3a is different from each other, and the second cylinder 3b. Although the ratio of the reciprocating mass and the rotating mass of the pair of crank webs 26b is different from each other, the ratio of the reciprocating mass and the rotating mass of the first cylinder 3a and the second cylinder 3b is formed to be equal to each other. Even if H is provided, the engine can be made compact without reducing the balancer function.

  As described above, the valve drive force transmission mechanism H that transmits power from the balancer shaft 32 to the valve mechanism G is provided, and power is not taken out from the crankshaft 24 to the valve mechanism G. An engine having a very good driving feeling can be made compact while ensuring driving of the driving force transmission mechanism H.

  Further, in this embodiment, as shown in FIGS. 3 and 4, the length L1 of the balancer shaft 32 is made smaller than the interval L2 between the crank webs 26a, 26d on both ends on the crankshaft 24, and one side in the crankshaft direction is set. The space K can be secured next to the balancer shaft 32 adjacent to the crankshaft 24 by the arrangement of the balancer shaft 32. A balance weight 35 at one end of the balancer shaft 32 is provided at a position corresponding to the crank web 26a of the cylinder 3a at one end, and a balance weight 36 at the other end is provided at the crank web 26c of the cylinder 3c adjacent to the cylinder 3d at the other end. The space K can be secured next to the balancer shaft 32 corresponding to the other end cylinder 3d. A power transmission part D is provided at the end of the balancer shaft 32 on the side 32a opposite to the side 32a that is displaced. The power transmission portion D is constituted by an output gear 38 formed integrally with the end portion of the balancer shaft 32.

  The crankcase 19 is provided with a shift mechanism 43 and a clutch mechanism 44, and the configuration of the shift mechanism 43 and the clutch mechanism 44 will be described with reference to FIG. The output gear 38 of the balancer shaft 32 is connected to the main shaft 41 via a multi-plate clutch mechanism 44. A multi-stage transmission gear 49 is mounted on the main shaft 41, and each transmission gear 49 is meshed with a transmission gear 420 mounted on the drive shaft 42 correspondingly. Either one or both of the transmission gear 49 and the transmission gear 420 are attached to the main shaft 41 or the drive shaft 42 in an idle state, except for the selected transmission gear. Therefore, rotation transmission from the main shaft 41 to the drive shaft 42 is performed only through the selected pair of transmission gears. It is transmitted from the drive shaft 42 to the rear wheel 13 via the rear wheel drive chain 13a.

  A shift operation for selecting the transmission gear 49 and the transmission gear 420 to change the transmission gear ratio is performed by a shift cam 421 that is a shift input shaft. The shift cam 421 has a plurality of cam grooves 421a, and a shift fork 422 is attached to each cam groove 421a. Each shift fork 422 is engaged with a predetermined transmission gear 49 and transmission gear 420 of the main shaft 41 and the drive shaft 42, respectively. Due to the rotation of the shift cam 421, the shift fork 422 moves in each axial direction according to the cam groove 421a, and only the pair of the transmission gear 49 and the transmission gear 420 at positions corresponding to the rotation angle of the shift cam 421 are the main shaft 41 and the drive shaft 42. On the other hand, both of them are fixed by the spline so that the rotation is transmitted and the transmission gear position is determined.

  The shift mechanism 43 rotates the shift cam 421 by a predetermined angle via the shift link mechanism 425 by driving the shift actuator 61. As a result, the shift fork 422 moves in the axial direction by a predetermined amount according to the cam groove 421a, and the pair of transmission gears 49 and the transmission gear 420 are sequentially fixed to the main shaft 41 and the drive shaft 42 to transmit the rotations of the respective reduction ratios. .

  The clutch mechanism 44 of this embodiment is, for example, a multi-plate friction clutch, and is rotatably provided with respect to the balancer shaft 32 that meshes with an output gear 38 that is integrally formed with the balancer shaft 32. An outer driver 443 that transmits torque from the balancer shaft 32 by being provided integrally with the gear 441, a plurality of friction disks 445 that are friction plates provided integrally with the outer driver 443, and an inner driver 447, which is a friction plate integrally provided on 447, and an inner member to which torque is transmitted from the outer driver 443 by frictional force generated between the plurality of friction disks 445 and the plurality of clutch plates 449. And a driver 447.

  The gear 441 is rotatably provided on the main shaft 41 on one end side of the main shaft 41, and the outer driver 443 is integrally provided on the boss portion of the gear 441, thereby rotating the rotation shaft of the main shaft 41. It is rotatable with respect to the main shaft 41 while being restricted from moving in the direction. The inner driver 447 is integrally provided on the main shaft 41 on one end portion side of the main shaft 41 (further end portion side than the gear 441).

  The inner driver 447 is provided inside the cylindrical outer driver 443, and the rotation centers of the gear 441, the outer driver 443, the inner driver 447, and the main shaft 41 are coincident and exist concentrically.

  An engagement portion having an engagement hole 443A engaged with a circular engagement protrusion 441A provided on a boss portion of the gear 441 on one end side of the opening of the cylindrical outer driver 443. 443B is provided, and the outer driver 443 is centered and fixed to the gear 441 using the engaging portion 443B and the engaging hole 443A. Each friction disk 445 is a ring-shaped thin plate, and the outer peripheral edge of each friction disk 445 is a cylinder so that the plane of each friction disk 445 is substantially perpendicular to the direction of the rotation axis of the main shaft 41. Is supported on the inner periphery of the cylindrical portion of the outer driver 443. With this support, each friction disk 445 is slightly movable relative to the outer driver 443 in the direction of the rotation axis of the main shaft 41, and the main shaft 41 with respect to the outer driver 443. It is restricted so that it cannot rotate relatively in the rotation direction.

  Further, a predetermined interval (a distance slightly larger than the thickness of the clutch plate 449) is provided between the planes of the friction disks 445.

  The inner driver 447 has a cylindrical shape, and is provided with a circular flange portion 447A having an outer diameter substantially equal to the outer diameter of the clutch plate 449 on one end side of the opening, and on the outer periphery of the cylindrical portion, A plurality of clutch plates 449 are supported. With this support, each clutch plate 449 is slightly movable relative to the inner driver 447 in the direction of the rotation axis of the main shaft 41, and the main shaft 41 with respect to the inner driver 447. It is restricted so that it cannot rotate relatively in the rotation direction.

  Further, the inner driver 447 is fixed to one end portion side of the main shaft 41 so that the flange portion 447A is on the engagement portion 443B side of the outer driver 443.

  Each clutch plate 449 is a ring-shaped thin plate, and the inner peripheral edge of each clutch plate 449 is a cylinder so that the plane of the plate of each clutch plate 449 is substantially perpendicular to the direction of the rotation axis of the main shaft 41. The inner driver 447 is supported on the outer periphery of the cylindrical portion.

  In addition, a predetermined distance (a distance slightly larger than the thickness of the friction disk 445) is provided between the planes of the clutch plates 449.

  The outer diameter of each clutch plate 449 is slightly smaller than the inner diameter of the cylindrical portion of the cylindrical outer driver 443, and the inner diameter of each friction disk 445 is smaller than the outer diameter of the cylindrical portion of the cylindrical inner driver 447. Slightly larger. The friction disks 445 and the clutch plates 449 are alternately arranged in the rotation axis direction of the main shaft 41. Between the friction disks 445 and the clutch plates 449, the rotation axis direction of the main shaft 41 is provided. There are slight gaps.

  Outer of each of the friction disks 445 and each of the clutch plates 449 arranged alternately and outside of the main shaft 41 in the rotation axis direction, on the side of the engaging portion 443B of the outer driver 443, the inner driver There is a pressing portion 447B configured by a flange portion 447A of 447. The pressing portion 447B sandwiches each friction disk 445 and each clutch plate 449 together with the pressure plate 451 in the rotation axis direction of the main shaft 41, and friction between each friction disk 445 and each clutch plate 449. It generates power.

  The pressure plate 451 is provided so as to be movable relative to the inner driver 447 in the direction of the rotation axis of the main shaft 41, and is configured to rotate simultaneously with the inner driver 447. The pressure plate 451 has a flat pressing portion 451B, and the pressing portion 451B is substantially parallel to the planes of the friction disks 445 and the clutch plates 449.

  The inner driver 447 has a plurality of cylindrical guide portions 447C for guiding the pressure plate 451, and a plurality of compression springs 450 are provided so as to surround each of the guide portions 447C. The spring 450 biases the pressure plate 451 in a direction in which the pressing portion 451B of the pressure plate 451 approaches the pressing portion 447B of the inner driver 447.

  Further, in a state where the clutch mechanism 44 is engaged and connected, the compression spring 450 moves the pressure plate 451 in the direction of the flange portion 447A of the inner driver 447 and is urged, and the pressure portion 447B of the inner driver 447 and the pressure. Each friction disk 445 and each clutch plate 449 are sandwiched and pressed by the pressing portion 451B of the plate 451, and a frictional force is generated between each friction disk 445 and each clutch plate 449. Torque can be transmitted to the inner driver 447.

  On the other hand, in a state where the clutch mechanism 44 is not engaged (a state in which the clutch mechanism 44 is not connected and torque is not transmitted), the pressure plate 451 is moved by a push rod 455 described later, and the pressing portion 451B of the pressure plate 451 is moved. It is separated from the friction disk 445 located closest to the pressing portion 451B of the pressure plate 451.

  Therefore, each friction disk 445 and each clutch plate 449 are not sandwiched, and there is a slight gap in the direction of the rotation axis of the main shaft 41 between each of these friction disks 445 and each clutch disk 449. A frictional force capable of transmitting torque is not generated between the plate 449 and the plate 449. Note that the pressure plate 451 is moved and controlled by the clutch actuator 60.

  The pressure plate 451 is moved in the direction of the rotation axis of the main shaft 41 by the clutch actuator 60 and the compression spring 450, and the clutch is engaged and engaged (becomes a state where power can be transmitted) according to this movement, or is not connected. It becomes a state (a state where power cannot be transmitted).

  The push rod 455 is attached to the compression spring 450 in the direction of the rotation axis of the main shaft 41 which is the left direction in FIG. 3 and the pressing portion 451B of the pressure plate 451 is separated from the pressing portion 447B of the inner driver 447. When moving with a force larger than the force, the pressure plate 451 is pushed by the push rod 455 and moves in the same manner.

  Conversely, when the push rod 455 moves to the left in FIG. 3, the pressure plate 451 moves in the same manner by pushing the push rod 455 with the urging force of the compression spring 450. A spherical ball 459 is provided inside the cylindrical main shaft 41 adjacent to the other end of the push rod 455, and a push rod 461 is provided adjacent to the ball 459. .

  The operation rod 461A1 is in contact with one end 461A of the push rod 461, and the operation rod 461A1 protrudes from the other end of the cylindrical main shaft 41. The operating rod 461A1 that protrudes from the push rod 461 is integrally provided with a piston 463 that constitutes the clutch actuator 60. The piston 463 is guided by the cylinder body 465 in the direction of the rotation axis of the main shaft 41. It is slidable.

  When hydraulic oil, which is an example of a compressed fluid, is supplied to a space 467 surrounded by the piston 463 and the cylinder body 465, the piston 463 is pushed to the right in FIG. 2 to move, and the operation rod 461A1, push The pressure plate 451 is pushed rightward in FIG. 3 via the rod 461, the ball 459, and the push rod 455. As described above, when the pressure plate 451 is pushed rightward in FIG. 3 and the pressing portion 451B of the pressure plate 451 is separated from the friction disk 445, the clutch is disconnected.

  Next, the case where the clutch mechanism 44 is changed from the non-connected state to the connected engagement state will be described. In a state where the clutch mechanism 44 is not connected, the piston 463 of the clutch actuator 60 pushes the pressure plate 451 rightward in FIG. 2 via the operation rod 461A1, the push rod 461, the ball 459, and the push rod 455. The pressing portion 451B of the pressure plate 451 is kept away from the friction disk 445. Even in this state, the pressure plate 451 is urged to the left in FIG. 3 by the compression spring 450, so that the piston 463 passes through the push rod 455, the ball 459, the push rod 461, and the operation rod 461A1. Thus, it is biased to the left in FIG.

  The piston 463 urged by the pressure plate 451 and the compression spring 450 gradually moves in the left direction in FIG. 3 due to the movement of the hydraulic oil from the state where the clutch mechanism 44 is not connected. Along with this, the pressure plate 451 gradually moves to the left in FIG. 3, and eventually the clutch mechanism 44 starts connection engagement (starts transmission of power), and the pressure plate 451 further moves to the left in FIG. As a result, the frictional force generated between the friction disk 445 and the clutch plate 449 increases due to the urging force of the compression spring 450, so that there is almost no slip between the friction disk 445 and the clutch plate 449, and the clutch engagement is achieved. Complete.

  Next, a support structure and a lubrication structure of the valve drive force transmission mechanism will be described with reference to FIGS. 9 is an exploded side view of the crankcase, FIG. 10 is a top view of the disassembled lower case, FIG. 11 is a top view of the disassembled upper case, FIG. 12 is a bottom view of the disassembled upper case, and FIG. FIG. 14 is a side view of the bearing cap, FIG. 15 is a cross-sectional view of the oil pump portion of the oil pan, FIG. 16 is a cross-sectional view of the lubricating oil supply passage from the oil pan to the bearing portion, 17 is a cross-sectional view taken along line XVII-XVII in FIG. 16, FIG. 18 is a side view of the cylinder body, FIG. 19 is a top view of the cylinder body, FIG. 20 is a cross-sectional view taken along line XX-XX in FIG. Is a sectional view taken along line XXI-XXI in FIG. 19, FIG. 22 is a sectional view taken along line XXII-XXII in FIG. 19, and FIG. 23 is a side view of the bearing cap.

  As shown in FIGS. 2, 3, and 5, the engine 1 includes a cylinder body 20 that forms a part of a cylinder having combustion chambers 80a to 80d at one end, and a part of the combustion chambers 80a to 80d. The cylinder head 21 to be joined to the cylinder body 20, the crank chambers 80a1 to 80d1 located at the other end of the cylinder are partly formed, the crank case 19 to be joined to the cylinder body 20, and the crank case 19 to be rotatably supported The crankshaft 24, the intake side camshaft 88a of the valve mechanism G that is rotatably supported by the cylinder head 21 above the combustion chambers 80a to 80d, and the camshaft that is the exhaust side camshaft 88b. Is provided with a valve drive force transmission mechanism H that transmits the rotation of the motor to the camshaft via the balancer shaft 32. In this embodiment, the rotation of the crankshaft 24 is transmitted to the camshaft via the balancer shaft 32, but without using the balancer shaft 32, the crankshaft 24 is directly transmitted via the valve drive force transmission mechanism H. It may be transmitted to the camshaft.

  A support structure for the valve drive force transmission mechanism by the joint surface of the cylinder body 20 and the crankcase 19 will be described with reference to FIGS. The crankcase 19 includes an upper case 17 and a lower case 18, and the upper case 17 has a lower joint surface 17a and an upper joint surface 17b, and the lower case 18 has an upper joint surface 18a. The lower joint surface 17a of the upper case 17 and the upper joint surface 18a of the lower case 18 are joined and fastened and fixed by the fastening bolt 100 and the nut 101. The crankshaft 24 is rotatably supported by five bearing support portions 19a1 to 19e1 formed in the upper case 17 and the lower case 18 via bearings 27a to 27e, respectively. The balancer shaft 32 is rotatably supported by four bearing support portions 19a to 19d formed in the upper case 17 and the lower case 18 via respective bearings 28a to 28d.

  The lower joint surface 20a of the cylinder body 20 shown in FIGS. 18 to 23 is joined to the upper joint surface 17b of the upper case 17 of the crankcase 19, and is fastened and fixed by the tightening bolts 110 and the nuts 111. As shown in FIGS. 12 to 14, a support portion 17c is formed on the upper joint surface 17b of the upper case 17 so as to extend in the center of the upper case 17 and to the side. The bearing 91a of the rotary shaft 90 is applied to the support portion 17c, and the bearing caps 112a and 112b shown in FIG. 13 are fastened and fixed by the mounting bolts 113 so that the rotary shaft 90 is pivotally supported by the bearing 91a. .

  The support portion 17c is provided with a bearing detent portion 17c1. The bearing detent portion 17c1 is formed in a hole shape. A locking pin 91a1 is formed on the bearing 91a. By engaging the locking pin 91a1 with the bearing detent portion 17c1 and assembling the bearing 91a to the support portion 17c, the bearing 91a is derotated to ensure lubrication. It is possible to improve the reliability.

  A support structure for the valve drive force transmission mechanism by the joint surface between the cylinder body 20 and the cylinder head 21 will be described with reference to FIGS. The cylinder body 20 has a lower joint surface 20a and an upper joint surface 20b. The cylinder body 20 is fastened and fixed to the upper joint surface 20b of the cylinder body 20 by being joined to the lower joint surface 21a of the cylinder head 21.

  A support portion 20c is formed on the upper joint surface 20b of the cylinder body 20 so as to extend in the center and to the side. The bearing 93a of the rotary shaft 92 is applied to the support portion 20c, and the bearing caps 122a and 122b shown in FIG. 23 are fastened and fixed by the mounting bolts 123 so that the rotary shaft 92 is pivotally supported by the bearing 93a. .

  The support portion 20c is provided with a bearing detent portion 20c1. The bearing detent portion 20c1 is formed in a hole shape. A locking pin 93a1 is formed on the bearing 93a. By engaging the locking pin 93a1 with the bearing detent portion 20c1 and assembling the bearing 93a to the support portion 20c, the bearing 93a is derotated to ensure lubrication. It is possible to improve the reliability.

  As described above, by using the joint surface between the cylinder body 20 and the cylinder head 21 and the joint surface between the cylinder body 20 and the crankcase 19 to support each part of the valve drive force transmission mechanism H, the compact and assemblability is achieved. The crankshaft direction width of the engine 1 can be made compact. Note that only the joint surface between the cylinder body 20 and the crankcase 19 may be used to support at least a part of the valve driving force transmission mechanism H, or only the joint surface between the cylinder body 20 and the cylinder head 21 may be used. Thus, at least a part of the valve driving force transmission mechanism H may be supported. In this way, it can be supported using the joint surface between the cylinder body 20 and the cylinder head 21 and the joint surface between the cylinder body 20 and the crankcase 19 without using a special support member for the valve drive force transmission mechanism H, and is compact. This is a straddle-type vehicle that is easy to assemble and that can make the crankshaft direction width of the engine 1 compact.

  Next, the lubricating oil supply path of the OHC engine will be described with reference to FIGS. An oil pan 52 is provided below the lower case 18 of the crankcase 19. The oil pan 52 stores lubricating oil for lubricating each lubricating portion of the engine 1, an oil strainer 53 is disposed inside, and an oil filter 54 is provided on the side portion.

  Further, as shown in FIG. 16, a lubricating oil return passage 52 a is formed in the side wall of the oil pan 52, and the lubricating oil falling through the return passage 18 m of the lower case 18 flows through the return passage 52 a of the oil pan 52. And returned to the inside of the oil pan 52. The lower case 18 is provided with an oil pump 120. The oil pump 120 rotates in conjunction with the crankshaft 24 by an interlocking mechanism (not shown), sucks the lubricating oil in the oil pan 52 through the oil strainer 53, It is supplied to a supply passage 18 c formed in the case 18. The lubricating oil flows from the supply passage 18c through the supply passage 52b formed on the side wall of the oil pan 52, passes through the oil filter 54, and further flows through the supply passage 52c to the supply passage 18c of the lower case 18.

  As shown in FIGS. 16 and 17, the supply passage 18c of the lower case 18 communicates with the supply passage 17e of the upper case 17, and the supply passage 17e communicates with a main gallery 17f formed in the crankshaft direction. . The branch supply passages 17f1 to 17f5 branch from the main gallery 17f, and lubricating oil flows through the branch supply passages 17f1 to 17f5 and is supplied to the bearings 27a to 27e of the bearing support portions 19a1 to 19e1 to be lubricated. Further, branch supply passages 17f21 and 17f22 are communicated with the branch supply passage 17f2, and lubricating oil flows through the branch supply passages 17f21 and 17f22 and is supplied to the crank pins 25a and 25b for lubrication. Branch supply passages 17f41 and 17f42 communicate with the branch supply passage 17f4, and lubricating oil flows through the branch supply passages 17f41 and 17f42 to be supplied to the crank pins 25c and 25d for lubrication.

  As shown in FIGS. 9 and 11, the main gallery 17 communicates with the sub gallery 17k via the transfer passage 17g. Branch passages 17k1 to 17k5 are formed from the sub gallery 17k, and the branch passages 17k1 to 17k5 open to the upper joint surface 17b of the upper case 17.

  As shown in FIG. 12, a supply passage 17 n communicates with the supply passage 17 e of the upper case 17, and the supply passage 17 n communicates with the supply passage 20 n of the cylinder body 20. Further, the branch passages 17k1 to 17k3 of the upper case 17 communicate with the supply passages 20k1 to 20k3 of the cylinder body 20. The supply passage 20n and the supply passages 20k1 to 20k3 of the cylinder body 20 are opened to the lower joint surface 20a and the upper joint surface 20b of the cylinder body 20.

  The supply passage 20k1 communicates with the supply passage 21k1 of the cylinder head 21 and supplies lubricating oil to the cylinder head 21 side. The supply passages 20k2 and 20k3 supply lubricating oil to the bearing 93a for lubrication.

  In this way, a part of the supply path of the lubricating oil supplied to the valve driving force transmission mechanism H is formed on the joint surface between the cylinder body 20 and the crankcase 19, and lubrication can be performed reliably. Reliability can also be improved, and a supply path can be easily formed from the joint surface. In addition, a part of the supply path of the lubricating oil supplied to the valve driving force transmission mechanism H is formed on the joint surface between the cylinder body 20 and the cylinder head 21, so that lubrication can be performed reliably and reliability. In addition, the supply path can be easily formed from the joint surface.

  Next, a configuration for eliminating the couple due to the primary inertia by the balancer shaft 32 will be described. 24 is a sectional view for explaining the operation of the crankshaft and the balancer, FIG. 25 is a perspective view schematically showing the crankshaft, FIG. 26 is a view showing the crank angle, and FIG. 27 is a view showing a change in the primary inertia force. FIG. 28 is a diagram schematically showing the configuration of the reciprocating portion, FIG. 29 is a perspective view for explaining the couple vector, FIG. 30 is a diagram for explaining the combined vector of the primary inertia force, and FIG. FIG. 3 is a perspective view showing a balance ratio of a crank web.

  The engine 1 of this embodiment is formed such that cylinder holes 3a1 to 3d1 for four cylinders are arranged in a line in the vehicle width direction, and the four cylinders 3a to 3d are arranged in order from the left side of the vehicle body, the first cylinder and the second cylinder. It is called 3rd cylinder and 4th cylinder.

  The balancer shaft 32 is for canceling the primary couple generated by the rotation of the crankshaft 24. The balancer shaft 32 has balance weights 35 and 36, and the rotation of the crankshaft 24 is coupled to the balancer shaft 32 and the clutch. Is transmitted to the main shaft 41 via the mechanism 44, and further transmitted from the main shaft 41 to the drive shaft 42 via the shift mechanism 43. The rotational direction of the crankshaft 24 is clockwise in FIG. It is.

  As shown in FIG. 26, the crank pin 25a to 25d in the rotational direction (crank angle) of the crankshaft 24 has a crank angle of 270 ° for the second cylinder crankpin 25b relative to the first cylinder crankpin 25a. The crank angle of the third cylinder crank pin 25c with respect to the first cylinder crank pin 25a is 90 °, and the crank angle of the fourth cylinder crank pin 25d with respect to the first cylinder crank pin 25a is 180 °. Has been. That is, in the crankshaft 24, the first cylinder crankpin 25a and the fourth cylinder crankpin 25d are positioned on the first imaginary plane (plane), and the second cylinder crankpin 25b and the third cylinder are used. The crank pin 25c is a two-plane, 180 ° type in which the two planes are perpendicular to each other and located on the second virtual plane. The ignition order of the engine 1 is as follows: No. 1 cylinder → No. 3 cylinder → No. 2 cylinder → No. 4 cylinder.

  The crank webs 26a to 26d are formed with counterweight portions on the opposite side across the axis of the crankshaft 24. The mass of the counterweight portion is set so that the exciting force of the engine 1 can be reduced in cooperation with the balancer shaft 32. By setting this mass to about 50% (the balance ratio is about 50%) of the mass of the reciprocating portion of the engine 1, the magnitude of the couple acting during rotation becomes appropriate. The couple having an appropriate magnitude here refers to a couple having a magnitude that can be canceled by the balancer shaft 32 that can be installed in the engine 1. The two-plane, 180 ° type crankshaft 24 according to this embodiment eliminates the primary, secondary inertial force and secondary couple, so the crank web is formed so that the balance ratio is about 50%. As a result, the movement locus of the couple synthesis vector of the remaining primary couple becomes a circle around the crankshaft center. That is, in the engine 1 equipped with the crankshaft 24, the primary couple can be eliminated by providing a couple balancer that generates a couple that becomes a couple vector that cancels the primary couple composite vector. it can. The reason why the primary and secondary inertia forces and the secondary couple are erased and the configuration of the couple balancer that erases the primary couple will be described below.

(1) About the reason why the primary inertial force is eliminated The primary inertial force acts on the crankshaft 24 so as to correspond to the reciprocation of the piston, as shown in FIG. In FIG. 27, the change in the primary inertia force due to the reciprocating mass of the first cylinder is shown by a solid line, the change of the primary inertia force due to the reciprocating mass of the second cylinder is shown by a broken line, and the reciprocating mass of the third cylinder is shown. The change of the primary inertia force due to the reciprocating mass of the fourth cylinder is indicated by a one-dot chain line. As shown in the figure, when the piston 22a of the first cylinder is located at the top dead center, the piston 22d of the fourth cylinder is located at the bottom dead center. Inertial forces cancel each other. Similarly, the primary inertia force due to the reciprocating mass of the second and third cylinders cancels each other out. Therefore, the primary inertia force is theoretically zero on the crankshaft 24.

(2) About the reason why the secondary inertial force is eliminated The secondary inertial forces F1 to F2 due to the reciprocating mass of each cylinder determine the dimensions and masses of the respective parts as shown in FIG. It is expressed as (4). In FIG. 28, mr = mass (g) of the reciprocating portion, L = connecting rod length (mm), r = piston stroke / 2 (mm), and ω = 2πN / 60 (rads). In the following formulas (1) to (4), λ = connecting rod length / r. F1 = mr × rω2 × 1 / λ × cos2θ (1)
F2 = mr × rω2 × 1 / λ × cos {2 (θ + 3π / 2} (2)
F3 = mr × rω2 × 1 / λ × cos {2 (θ + π / 2} (3)
F4 = mr × rω2 × 1 / λ × cos {2 (θ + π)} (4)
Since the sum F (2) of F1 to F4 is 0 as shown below, the secondary inertia force is theoretically 0 on the crankshaft 24.

Number 1

(3) About the reason why the secondary couple is eliminated The secondary couple is the sum of moments around the Y axis generated by the secondary inertia forces F1 to F4 acting on the crankshaft 24, as shown in FIG. It is. The Y axis is an axis that is orthogonal to the X axis parallel to the cylinder axis and extends in a direction orthogonal to the Z axis parallel to the crankshaft 24 axis. FIG. 25 shows a state where the piston 22a of the first cylinder is located at the top dead center. Since the secondary inertia forces F1 to F4 are repeatedly generated so that one cycle has a crank angle of 180 °, the first and fourth cylinders having different crank angles of 180 ° have the same direction in which the secondary inertia force acts. Become a direction. The second cylinder in which the crank angle is shifted by 90 ° to the front side in the rotational direction with respect to the first cylinder and the third cylinder in which the crank angle is shifted by 90 ° to the rear side in the rotational direction are 1, 4 in the direction in which the secondary inertia force acts. The direction is the opposite of the numbered cylinder. Therefore, the secondary couple FL consisting of the sum of the moments becomes 0 as shown below when the distance from the Y axis to each cylinder is d1 to d4 and d1 = d4 and d2 = d3.

FL = F1 * d1-F2 * d2 + F3 * d3-F4 * d4 = 0
For this reason, the secondary couple is zero on the crankshaft 3.

(4) About the structure of the couple balancer 6 that eliminates the remaining primary couple The primary couple acting on the crankshaft 24 is represented by the sum of vectors of primary inertial forces acting on the crank pins 25a to 25d. Can do. This will be described in detail with reference to FIG. FIG. 29 shows a state in which the piston 22a of the first cylinder is located at the top dead center, that is, a state in which the first cylinder crank pin 25a is located on the X axis (cylinder axis) and at the upper end in FIG. Is drawn.

  At this time, in the first cylinder, the primary inertia force due to the mass of the reciprocating portion (piston 22a, piston pin 25a and connecting rod 23a) and the rotating portion (crank pin 25a, connecting rod 23a large end portion and The resultant force with the primary inertia force due to the mass of the crank webs 26a, 26a) acts upward. This is because the crankshaft 24 is formed so as to have a balance of about 50% as described above, so that the inertial force acting toward the upper part of the reciprocating motion part is directed downward of the rotational motion part. This is because the inertial force acting on is reduced by half. That is, in the first cylinder, the resultant vector of the primary inertial force due to the reciprocating mass and the rotating mass is directed upward in parallel with the X axis. The magnitude of the composite vector of the first cylinder changes as shown in FIG. The direction of rotation of the combined vector is counterclockwise opposite to the direction of rotation of the crankshaft.

  In the second cylinder, since the piston 25b is located at a substantially intermediate point in the middle of moving from the bottom dead center to the top dead center, the inertial force of the reciprocating motion portion is substantially zero. For this reason, in the second cylinder, only the inertial force of the rotating part (acts so as to be directed toward the rear of the vehicle body) acts, so the resultant vector of the primary inertial force is rearward in parallel with the Y axis (see FIG. In the case of 25, it is directed to the right). The magnitude of the composite vector of the second cylinder changes as shown in FIG. The direction of rotation of the combined vector is counterclockwise opposite to the direction of rotation of the crankshaft.

  In the third cylinder, since the piston 25c is located at a substantially middle point in the middle of moving from the top dead center to the bottom dead center, the inertial force of the reciprocating motion portion is substantially zero, and only the inertial force of the rotating portion acts. For this reason, the resultant vector of the primary inertia force of the third cylinder is directed forward (to the left in FIG. 29) parallel to the Y axis. The size of the composite vector of the third cylinder changes as shown in FIG. The direction of rotation of the combined vector is counterclockwise opposite to the direction of rotation of the crankshaft.

  In the fourth cylinder, since the piston 25d is located at the bottom dead center, the inertial force of the reciprocating motion part is directed downward, and the inertial force of the rotational motion part, which is about half of the reciprocating motion part, is directed upward. To do. For this reason, the resultant vector of the primary inertia force of the fourth cylinder is directed downward in parallel with the X axis. The magnitude of the composite vector of the fourth cylinder changes as shown in FIG. The direction of rotation of the combined vector is counterclockwise opposite to the direction of rotation of the crankshaft.

  The combined vector of the first cylinder and the combined vector of the fourth cylinder are 180 ° opposite to each other and have substantially the same magnitude, so the primary inertia force acting on the crankpins 25a and 25d of these two cylinders is a couple. It becomes. This couple is hereinafter referred to as the first couple. This first couple is a couple that turns the crankshaft 24 around a virtual axis YC that passes through the center of the crankshaft 24 in the axial direction and is parallel to the Y axis in FIG. The vector of the first couple is the direction in which the right screw advances when the couple is in the same direction as the rotation direction when the right screw is tightened. That is, the first couple vector is directed forward along the virtual axis YC. In FIG. 29, a plane on which the virtual axis YC is located and perpendicular to the axis of the crankshaft 24 is drawn as a virtual plane I at the left end of the same figure for easy understanding. On the virtual plane I, the vector of the first couple is denoted by A.

  On the other hand, the combined vector of the second cylinder and the combined vector of the third cylinder are 180 degrees opposite to each other and have substantially the same magnitude, so that the second inertial force acting on the crankpins of these two cylinders causes the second Couples are generated. Since the second couple vector is the direction in which the right screw advances when the couple is in the same direction as the rotation direction when the right screw is tightened, the center in the axial direction of the crankshaft 3 in FIG. The direction is directed upward along a virtual axis XC parallel to the X axis. On the virtual plane I, the vector of the second couple is denoted by B.

  By combining these vectors A and B, the direction and magnitude of the total couple acting on the engine 1 can be obtained. The combined vector T of the vector A and the vector B is directed diagonally upward and forward as shown on the virtual plane I in FIG. The magnitude of this combined vector changes as shown in FIG. The rotation direction of this combined vector is counterclockwise, which is the opposite of the rotation direction of the crankshaft.

  Since this composite vector T represents all the primary couples of the engine 1, the couple of a vector (indicated by a broken line t in FIG. 29) that is point-symmetric with this composite vector T is expressed by the balancer shaft 32. By generating, all the primary couples generated in the engine 1 can be eliminated. A vector t that is point-symmetric with the composite vector T is hereinafter referred to as a balancer vector.

  In order to balance the synthesized vector T and the balancer vector t, it is most effective that the rotation locus of the synthesized vector T shown in FIG. In this embodiment, the rotational trajectory is set to a substantially perfect circle by setting the weight balance of the crank web to about 50%. In the balancer shaft 32 having the balancer vector t, an inertial force is generated in a direction indicated by broken lines br1 and br2 in FIG. That is, by forming the balancer shaft 32 so that the inertial force becomes br1 and br2, the primary couple of the engine 1 can be balanced by the balancer shaft 32 and eliminated.

  Here, a method of determining the weight balance of the crank web so that the rotation locus becomes a circle will be described. The reciprocating mass of each cylinder is Mr (g), the rotational mass of each cylinder is Mo (g), and the unbalance amount (moment: gmm) of two crank webs per cylinder is sequentially from M1 to M8. Then, the balance ratio k of each crank web is k = {M (1-8) −Mo / 2} / Mr.

  The unbalance amounts M1 to M8 for making the combined vector of the primary inertia force constant (circle) in all the cylinders are M1 to M8 = 0.25 × Mr + Mo / 2. If M1 and M4 are equal and M2 and M3 are equal, one balance ratio k of the pair of crank webs is formed to be smaller than 0.25, and the other balance ratio k is set accordingly. It can be formed to be only large.

  In this embodiment, the balance ratio of each crank web is set as shown in FIG. That is, for the first and fourth cylinders, the balance ratio of the crank webs 26a and 26d located outside the engine is 0.427, and the balance ratio of the crank webs 26a and 26d located inside the engine 1 is 0.025. , Both crank webs were formed to have a balance ratio of 0.452. For the second and third cylinders, the balance ratio of the crank webs 26b and 26c located outside the engine 1 is 0.357, and the balance ratio of the crank webs 26b and 26c located inside the engine 1 is 0.017. As a result, the balance ratio of both crank webs was 0.374.

  As shown in FIGS. 2 and 24, the balancer shaft 32 is disposed at a position spaced rearward from the crankshaft 24 (rearward along the Y axis shown in FIG. 29), and the first and second balance weights 35 are disposed. , 36 are provided at positions corresponding to the crank pins 25b, 25c of the second and third cylinders. The first and second balance weights 35 and 36 according to this embodiment are formed so as to have a fan shape when viewed from the axial direction, and the second cylinder crank webs 26b and 26b are moved to the most crankshaft side. And between the third cylinder crank webs 26c and 26c. The balance weights 35 and 36 of the balancer shaft 32 are embedded with heavy metal in order to adjust the mass. The balance weight 35 generates an inertial force indicated by br1 in FIG. 29, and the balance weight 36 is formed so as to generate an inertial force indicated by br2 in FIG.

  As shown in FIGS. 24 (a) to 24 (e), the balancer shaft 32 according to this embodiment is as close to the crankshaft 24 as possible while avoiding contact between the balance weights 35 and 36 and the large end portion 23a1b of the connecting rod 23a. It is arranged at a close position. FIG. 24A shows the positions of the crank pin 25a of the first cylinder and the balance weight 35 when the piston 22a of the first cylinder is located at the top dead center, and FIG. The state where the crankshaft 24 is rotated 180 ° from the state shown in the drawing is shown. FIG. 4C shows a state where the crankshaft 24 has rotated 217.5 ° from the state of FIG. 5A, and FIG. 4D shows the state where the crankshaft 24 has rotated 225 ° from the state of FIG. The state (e) shows a state where the crankshaft 24 has rotated 270 ° from the state shown in FIG. Note that the positions of the large end portion 23c1 of the third cylinder connecting rod 23c and the balance weight 36 are the same as those in FIG. The in-line four-cylinder engine 1 configured as described above can erase the primary inertia force, the secondary inertia force, and the secondary couple by only the crankshaft 24, and the remaining primary couple is transferred to the balancer shaft. 32 can be erased.

  In the present invention, the length of the balancer shaft is made smaller than the interval between the crank webs at both ends on the crankshaft, and is deviated to one side in the crankshaft direction. Can be secured. By arranging a power transmission unit, a clutch, and the like in this space, the entire engine can be made compact while the driving feeling is very good.

1 is a side view of a motorcycle. It is sectional drawing of a parallel multicylinder 4 cycle engine. It is sectional drawing which shows arrangement | positioning of a crankshaft, a balancer shaft, and a main shaft. It is sectional drawing which shows arrangement | positioning of a crankshaft and a balancer shaft. It is a side view which shows a valve drive force transmission mechanism. It is a side view which shows the cam shaft part of a valve mechanism. It is a side view which shows arrangement | positioning of a crankshaft and a balancer shaft. It is the figure which looked at the cam shaft of the valve operating mechanism from the cam shaft direction. It is the side view which disassembled the crankcase. It is a top view of the decomposed lower case. It is a top view of the disassembled upper case. It is a bottom view of the disassembled upper case. It is sectional drawing which follows the XIII-XIII line | wire of FIG. It is a side view of a bearing cap. It is sectional drawing of the oil pump part of an oil pan. It is sectional drawing of the supply path of the lubricating oil from an oil pan to a bearing part. It is sectional drawing which follows the XVII-XVII line of FIG. It is a side view of a cylinder body. It is a top view of a cylinder body. It is sectional drawing which follows the XX-XX line of FIG. It is sectional drawing which follows the XXI-XXI line | wire of FIG. It is sectional drawing which follows the XXII-XXII line | wire of FIG. It is a side view of a bearing cap. It is sectional drawing for demonstrating operation | movement of a crankshaft and a balancer. It is a perspective view which shows a crankshaft typically. It is a figure which shows a crank angle. It is a graph which shows the change of a primary inertia force. It is a figure which shows typically the structure of a reciprocating motion part. It is a perspective view for demonstrating a couple vector. It is a figure for demonstrating the synthetic | combination vector of a primary inertia force. It is a perspective view which shows the balance rate of a crank web.

Explanation of symbols

1 Engine 3a-3d Cylinder 3
19 Crankcase 19a-19d Four bearing support portions of crankcase 19 24 Crankshaft 27a-27e Crankshaft 24 bearing 28a-28d Balancer shaft 32 bearing 28a-28d
32 Balancer shaft 35, 36 Balance weight

Claims (9)

A crankshaft having a crankpin that supports a connecting rod provided for each cylinder, and a crank web that supports the crankpin;
A parallel multi-cylinder four-cycle engine that is arranged in parallel to the crankshaft and has a balance weight and eliminates a couple due to primary inertia, wherein the length of the balancer shaft is set on the crankshaft. Smaller than the distance between the crank webs at both ends of the
A parallel multi-cylinder four-cycle engine characterized in that the balancer shaft is arranged to be displaced toward one side in the crankshaft direction. A crankshaft having a crankpin that supports a connecting rod provided for each cylinder, and a crank web that supports the crankpin;
A parallel multi-cylinder four-cycle engine that is arranged in parallel with the crankshaft and has a balance weight and eliminates the couple due to primary inertia;
The balancer shaft has a plurality of balance weights,
The balance weight at one end is provided at a position corresponding to the crank web of the cylinder at one end,
A parallel multi-cylinder four-cycle engine characterized in that the balance weight at the other end is disposed at a position corresponding to a crank web of a cylinder adjacent to the cylinder at the other end.   The balancer shaft is supported so that a rotation locus of the balance weight provided on the balancer shaft and a rotation locus of the crank web are overlapped with each other when viewed in the crankshaft direction. The parallel multi-cylinder four-cycle engine described.   4. The parallel multi-cylinder four-cycle engine according to claim 2, wherein a power input unit for inputting power extracted from the crankshaft is provided in an intermediate portion between balance weights of the balancer shaft. .   The power transmission part which transmits the motive power from a balancer axis | shaft to the exterior was provided in the edge part on the opposite side to the side to which the balancer axis | shaft was displaced. The parallel multi-cylinder four-cycle engine described in 1.   The parallel multi-cylinder four-cycle engine according to any one of claims 1 to 5, wherein the balancer shaft bearings are arranged corresponding to the crankshaft bearings.   The parallel multi-cylinder four-cycle engine according to claim 6, wherein the bearing of the balancer shaft is disposed between the balance weight and the power take-out portion.   The parallel multi-cylinder four-cycle engine according to claim 6, wherein bearings of the balancer shaft are arranged on both sides of the power take-out portion.   A straddle-type vehicle equipped with the parallel multi-cylinder four-cycle engine according to any one of claims 1 to 9.

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