JP2015007406A - Engine and vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide an engine and a vehicle that can reduce hitting sounds of a crankshaft while restraining an increase in the size of a crankcase, complication of its structure, and an increase in its weight.SOLUTION: A power transmission member that transmits a torque for driving a rear wheel 7 is attached to a crankshaft. The crankshaft is supported by bearings 230 and 240 of the same kind. The bearing 240 is located farther from the power transmission member than the bearing 230. A radial bearing clearance of the bearing 240 is smaller than a radial bearing clearance of the bearing 230.

Description

  The present invention relates to an engine and a vehicle including a bearing that supports a crankshaft.

  In an engine of a vehicle such as a motorcycle, a crankshaft is rotatably supported by a crankcase by a plurality of bearings. A crankpin of the crankshaft is connected to the piston via a connecting rod (connecting rod). The reciprocating motion of the piston is converted into the rotational motion of the crankshaft. At this time, the pressure from the piston is transmitted to the crankshaft through the connecting rod. As a result, a hitting sound may occur due to rattling of the crankshaft. Therefore, it is required to suppress the hitting sound (for example, Patent Document 1).

  In the bearing structure described in Patent Document 1, the left journal shaft portion of the crankshaft is supported by a ball bearing, and the right journal shaft portion is supported by a roller bearing. A gear group is fixed to the right end of the crankshaft. The outer race (outer ring) of the ball bearing is loosely fitted into one bearing hole of the crankcase, and the outer race of the roller bearing is press-fitted into the other bearing hole. This facilitates assembly and maintenance of the crankshaft to the crankcase while suppressing backlash of the crankshaft in the roller bearing.

JP 2005-105921 A

  In patent document 1, in addition to said structure, the backlash absorption means is provided. In the backlash absorbing means, the push plug is disposed between the cylinder and the crankshaft in parallel with the crankshaft. The push plug is biased so as to press the outer race of the ball bearing against the side receiving the explosion pressure in the bearing hole of the crankcase. That is, the outer race of the ball bearing is pressed in a direction perpendicular to the axis of the crankshaft by the push plug. This prevents backlash of the crankshaft in the ball bearing.

  However, in order to provide the backlash absorbing means in the crankcase, it is necessary to greatly change the design of the crankcase. In addition, the crankcase becomes larger and the structure of the crankcase becomes complicated and the weight increases.

  An object of the present invention is to provide an engine and a vehicle capable of reducing the hitting sound of a crankshaft while suppressing an increase in the size of a crankcase, a complicated structure, and an increase in weight.

  (1) An engine according to a first aspect of the present invention is an engine that generates torque for rotating a wheel, and includes a crankshaft to which a power transmission member that transmits torque for driving the wheel is attached, and a crankshaft First and second bearings of the same type to be supported, the second bearing is located farther from the power transmission member than the first bearing, and the radial bearing clearance of the second bearing is This is smaller than the bearing gap in the radial direction of No. 1 bearing.

  In this engine, a power transmission member that transmits torque for driving wheels is attached to a crankshaft. The crankshaft is supported by the same type of first and second bearings. The second bearing is located farther from the power transmission member than the first bearing.

  Under such a configuration, the inventor of the present application has found that there is the following mechanism regarding the hitting sound of the crankshaft.

  When the crankshaft transmits torque via the power transmission member, the crankshaft is slightly inclined by receiving a reaction force from the power transmission member. In this case, the distance between the second bearing and the power transmission member is longer than the distance between the first bearing and the power transmission member.

  Therefore, when the crankshaft receives an impact load due to the explosion pressure from the piston, a radial gap (hereinafter referred to as a moving gap) in which a portion supported by the second bearing of the crankshaft moves is referred to as a crankshaft. This is larger than the radial movement gap of the portion supported by the first bearing. Here, the moving clearance when the rolling bearing is used is the sum of the clearance between the outer peripheral surface of the crankshaft and the inner peripheral surface of the bearing and the internal clearance of the bearing. Further, the moving gap in the case of using the sliding bearing is a gap between the outer peripheral surface of the crankshaft and the inner peripheral surface of the bearing. In such a situation, the inventor of the present application, when subjected to an impact load due to the explosion pressure, is affected by the impact of the sound generated by the movement of the portion supported by the second bearing of the crankshaft through the moving gap. It was found that the effect of this part is greater than the impact on the sound.

  According to said structure, the radial bearing clearance of a 2nd bearing is smaller than the radial bearing clearance of a 1st bearing. For this reason, even when the crankshaft receives a reaction force from the power transmission member, the radial movement gap of the portion of the crankshaft supported by the second bearing is reduced. Thereby, the hitting sound due to the radial displacement of the crankshaft is reduced. In this case, it is not necessary to provide an additional structure in the crankcase in order to reduce the amount of displacement of the crankshaft. As a result, it is possible to reduce the hitting sound of the crankshaft while suppressing an increase in the size of the crankcase of the engine, a complicated structure, and an increase in weight.

  (2) The first and second bearings are first and second rolling bearings, respectively, the radial bearing gap of the first bearing is a radial internal gap of the first rolling bearing, and the second bearing The radial bearing gap may be a radial internal gap of the second rolling bearing.

  In this case, the use of rolling bearings as the first and second bearings restricts the movement of the crankshaft in the thrust direction. Thereby, the hitting sound due to the displacement of the crankshaft in the thrust direction is reduced. As a result, the hitting sound of the crankshaft can be reduced with a simple configuration.

  (3) In a state where the crankshaft and the second bearing are assembled as an engine, the radial internal clearance of the second bearing may be 0 or less.

  In this case, the radial movement clearance of the portion supported by the second bearing of the crankshaft is reduced. Thereby, the hitting sound of the crankshaft can be further reduced. Since each of the first and second bearings is a rolling bearing, the rotation of the crankshaft is not hindered even if the radial internal clearance of the second bearing is 0 or less. Further, when the radial internal clearance of the second bearing is 0 or less, the hitting sound of the crankshaft is sufficiently reduced. Therefore, the radial internal clearance of the first bearing need not be 0 or less. Therefore, an increase in friction torque can be minimized.

  (4) The inner ring of the second bearing may be attached to the crankshaft by an interference fit.

  In this case, the play of the crankshaft in the radial direction in the inner ring of the second bearing can be prevented. Thereby, the hitting sound of the crankshaft can be reduced.

  (5) The engine may further include a crankcase that holds the first and second bearings, and the outer ring of the second bearing may be attached to the crankcase by an interference fit.

  In this case, rattling of the second bearing in the crankcase is prevented. Further, the radial bearing gap of the second bearing can be easily made smaller than the radial bearing gap of the first bearing. Thereby, the hitting sound of the crankshaft can be reduced.

  (6) The first and second bearings are first and second sliding bearings, respectively, and the radial bearing clearance of the first bearing is between the inner surface of the first sliding bearing and the outer peripheral surface of the crankshaft. The radial bearing clearance of the second bearing may be a clearance between the inner surface of the second sliding bearing and the outer peripheral surface of the crankshaft.

  Sliding bearings have higher rigidity than other types of bearings. Therefore, by using the first and second bearings as the first and second sliding bearings, respectively, it is possible to reduce vibration in the frequency range where the crankshaft sound is generated. Thereby, the hitting sound of the crankshaft can be reduced with a simple configuration.

  (7) The engine further includes first and second housings that hold the first and second bearings, and a crankcase that holds the first and second housings. It may be attached to the housing by an interference fit.

  In this case, rattling of the second bearing in the crankcase is prevented. Further, the radial bearing gap of the second bearing can be easily made smaller than the radial bearing gap of the first bearing. Thereby, the hitting sound of the crankshaft can be reduced.

  (8) The rigidity of the second bearing may be higher than the rigidity of the first bearing.

  In this case, the vibration of the portion of the crankshaft supported by the second bearing is reduced. Thereby, the hitting sound of the part supported by the 2nd bearing of a crankshaft can be reduced more.

  (9) The rigidity of the first bearing may be higher than the rigidity of the second bearing. In this case, the vibration of the portion of the crankshaft supported by the first bearing is reduced.

  (10) The power transmission member may be a belt of a belt type continuously variable transmission. According to this configuration, it is possible to reduce the hitting sound of the crankshaft even when the belt pulls the crankshaft.

  (11) A vehicle according to a second aspect of the invention includes the engine according to the first aspect of the invention and a main body that moves by torque generated by the engine.

  In this vehicle, the main body moves by torque generated by the engine. In the engine, even when the crankshaft transmits torque through the power transmission member, the radial movement clearance of the portion of the crankshaft supported by the second bearing is reduced. Thereby, the hitting sound due to the radial displacement of the crankshaft is reduced. In this case, it is not necessary to provide an additional structure in the crankcase in order to reduce the amount of displacement of the crankshaft. As a result, it is possible to reduce the hitting sound of the crankshaft while suppressing an increase in the size of the crankcase of the engine, a complicated structure, and an increase in weight.

  According to the present invention, it is possible to reduce the hitting sound of the crankshaft while suppressing an increase in the size of the crankcase, a complicated structure, and an increase in weight.

1 is a schematic side view showing a schematic configuration of a motorcycle according to an embodiment of the present invention. It is a horizontal sectional view of a power transmission device seen from the upper part. It is an expanded sectional view of a bearing. It is an expanded sectional view of a bearing and its periphery. It is the sectional view on the AA line of FIG. It is an expanded sectional view of a crank web and its periphery. It is a schematic diagram which shows the other example of the combination of a bearing. It is a schematic diagram which shows the further another example of the combination of a bearing. It is a horizontal sectional view of a power transmission device seen from the upper part in case a bearing is a slide bearing. It is an expanded sectional view of a bearing in case a bearing is a slide bearing. It is a schematic diagram which shows the relationship between a combustion load direction and the direction of the tensile force by a V belt. It is a schematic diagram which shows the structure of an engine at the time of using a gear group as a power transmission member.

(1) Motorcycle FIG. 1 is a schematic side view showing a schematic configuration of a motorcycle according to an embodiment of the present invention. The motorcycle shown in FIG. 1 is an example of a vehicle. In the following description, front, rear, left and right mean front, rear, left and right based on the viewpoint of the motorcycle driver. In the motorcycle 100 of FIG. 1, a front fork 2 is provided at the front portion of the vehicle body 1 so as to be swingable in the left-right direction. A handle 4 is attached to the upper end of the front fork 2, and a front wheel 3 is rotatably attached to the lower end of the front fork 2.

  A seat 5 is provided at a substantially upper center portion of the vehicle body 1. A control device 6 is provided below the seat 5. In the present embodiment, the control device 6 is, for example, an ECU (Electronic Control Unit). A power transmission device 500 is provided below the control device 6. The power transmission device 500 includes, for example, a 4-stroke single cylinder engine 200 (see FIG. 2 described later). A rear wheel 7 is rotatably attached to the lower rear end of the vehicle body 1. The rear wheel 7 is rotationally driven by the power generated by the engine 200 of the power transmission device 500.

  FIG. 2 is a horizontal sectional view of the power transmission device 500 as viewed from above. As shown in FIG. 2, power transmission device 500 includes an engine 200 and a V-belt continuously variable transmission (CVT) 300. Engine 200 includes a crankcase 210, a crankshaft 220, bearings 230 and 240, a cylinder 250, a piston 260, and a connecting rod (connecting rod) 270.

  In this example, the crankcase 210 is a left-right split crankcase configured by fitting a left case member 211 and a right case member 212 together. The left case member 211 and the right case member 212 are made of aluminum, for example. The bearings 230 and 240 are supported by the left case member 211 and the right case member 212, respectively.

  In this example, the crankshaft 220 is an assembled crankshaft and includes a pair of crank journals 221 and 222, a pair of crank webs 223 and 224, and a crankpin 225. Each of the crank webs 223, 224 includes a crank arm and a balance weight. In this example, the crank journal 221 and the crank web 223 are integrally formed, and the crank journal 222 and the crank web 224 are integrally formed.

  The large end of the connecting rod 270 is disposed between the crank arms of the crank webs 223 and 224. In this state, the crank arms of the crank webs 223 and 224 and the large end of the connecting rod 270 are connected to each other by the crank pin 225 so as to be rotatable. The left crank journal 221 is rotatably supported by the bearing 230, and the right crank journal 222 is rotatably supported by the bearing 240. The piston 260 is provided so as to reciprocate within the cylinder 250. The small end of the connecting rod 270 is connected to the piston 260 by a piston pin.

  CVT 300 includes a transmission case 310, a drive pulley 320, a driven pulley 330, and a V belt 340. The transmission case 310 houses the drive pulley 320, the driven pulley 330, and the V belt 340, and rotatably supports the main shaft 120 and the rear wheel shaft 140 of the motorcycle 100. The drive pulley 320 is attached to the left end portion of the crankshaft 220. Accordingly, the bearing 240 is located on the opposite side of the V-belt 340 with respect to the bearing 230. The driven pulley 330 is attached to the main shaft 120. The V belt 340 is stretched over the drive pulley 320 and the driven pulley 330.

  The reciprocating motion of the piston 260 is converted into the rotational motion of the crankshaft 220 via the connecting rod 270. The rotational movement of the crankshaft 220 is transmitted to the driven pulley 330 via the drive pulley 320 and the V belt 340. The rotational motion of the driven pulley 330 is transmitted to the main shaft 120 via the start clutch 110. The rotational movement of the main shaft 120 is transmitted to the rear wheel shaft 140 via the gear mechanism 130. As the rear wheel shaft 140 rotates, the rear wheel 7 fixed to the rear wheel shaft 140 rotates.

(2) Engine Configuration FIG. 3 is an enlarged cross-sectional view of the bearings 230 and 240. 3A and 3B show bearings 230 and 240, respectively. The bearings 230 and 240 are of the same type as each other. In this example, the bearings 230 and 240 are rolling bearings. Rolling bearings include ball bearings and roller bearings. The roller bearing includes a cylindrical roller bearing, a rod roller bearing, a needle roller bearing, and a tapered roller bearing. The bearings 230 and 240 restrict the movement of the crankshaft 220 in the thrust direction (the axial direction of the crankshaft 220). Thereby, the hitting sound due to the displacement of the crankshaft 220 in the thrust direction is reduced.

  As shown in FIG. 3A, the bearing 230 includes an inner ring (inner race) 231, an outer ring (outer race) 232, a rolling member 233 and a holder (retainer) 234 (see FIG. 5 described later). The inner ring 231 and the outer ring 232 have an annular shape. Thereby, a substantially circular hole 230h is formed at the center of the inner ring 231. The plurality of rolling members 233 are arranged at substantially equal intervals between the inner ring 231 and the outer ring 232 by the holder 234.

  In this example, the bearing 230 which is a rolling bearing is a ball bearing. Therefore, the rolling member 233 has a spherical shape. A gap between the rolling member 233 and the inner ring 231 is called a gap 231c, and a gap between the rolling member 233 and the outer ring 232 is called a gap 232c. The sum of the gap 231c and the gap 232c, that is, the radial internal gap of the bearing 230 becomes the bearing gap 230c in the radial direction of the bearing 230 (radial direction of the crankshaft 220).

  Similarly, as shown in FIG. 3B, the bearing 240 includes an inner ring 241, an outer ring 242, a rolling member 243, and a holder (not shown). The inner ring 241 and the outer ring 242 have an annular shape. Thereby, a substantially circular hole 240 h is formed at the center of the inner ring 241. The plurality of rolling members 243 are arranged at substantially equal intervals between the inner ring 241 and the outer ring 242 by a holder.

  In this example, the bearing 240 which is a rolling bearing is a ball bearing. Therefore, the rolling member 243 has a spherical shape. A gap between the rolling member 243 and the inner ring 241 is referred to as a gap 241c, and a gap between the rolling member 243 and the outer ring 242 is referred to as a gap 242c. The sum of the gap 241 c and the gap 242 c, that is, the radial internal gap of the bearing 240 becomes the bearing gap 240 c in the radial direction of the bearing 240.

  As shown in FIGS. 3A and 3B, the radial bearing gap 240 c of the bearing 240 is smaller than the radial bearing gap 230 c of the bearing 230. In this case, the radial movement gap of the crank journal 222 supported by the bearing 240 can be reduced.

  Here, the movement clearance of the crank journal 221 includes a clearance between the outer peripheral surface of the crank journal 221 and the inner peripheral surface of the bearing 230, a radial internal clearance of the bearing 230, and a hole described later of the outer peripheral surface of the bearing 230 and the left case member 211. 211h (a hole 213h to be described later in this example) is a total of gaps with the inner peripheral surface. The movement clearance of the crank journal 222 includes a clearance between the outer peripheral surface of the crank journal 222 and the inner peripheral surface of the bearing 240, a radial internal clearance of the bearing 240, and a hole 212h (this) of the outer peripheral surface of the bearing 240 and the right case member 212 described later. In the example, it is the total of the gaps with the inner peripheral surface of the holes 213h) described later.

  The crank journal 221 of the crankshaft 220 is inserted into the hole 230 h of the bearing 230. Further, the crank journal 222 of the crankshaft 220 is inserted into the hole 240 h of the bearing 240. Thereby, the crankshaft 220 is supported by the bearings 230 and 240. The inner ring 241 of the bearing 240 is preferably attached to the crankshaft 220 by an interference fit. In this example, the crank journal 222 is press-fitted into the hole 240 h of the bearing 240. Thereby, the radial play of the crank journal 222 in the inner ring 241 of the bearing 240 is prevented.

  In a state where the crankshaft 220 and the bearing 240 are assembled to the crankcase 210, the bearing clearance 240c in the radial direction of the bearing 240 is preferably 0 or less (0 or negative). Since each of the bearings 230 and 240 is a rolling bearing, the rotation of the crankshaft 220 is not hindered even if the bearing gap 240c of the bearing 240 is 0 or less. Further, as will be described later, the hitting sound of the crankshaft 220 is sufficiently reduced, so that it is not necessary to reduce the bearing gap 230c of the bearing 230 to 0 or less. Therefore, an increase in friction torque can be minimized.

  In this example, the outer diameter of the bearing 230 is larger than the outer diameter of the bearing 240, and the rigidity of the bearing 230 is higher than the rigidity of the bearing 240. Here, the rigidity is expressed by a rated load. In this case, the vibration of the crank journal 221 supported by the bearing 230 is reduced.

  FIG. 4 is an enlarged cross-sectional view of the bearings 230 and 240 and the periphery thereof. As shown in FIG. 4, the crankcase 210 has a pair of inserts 213 made of, for example, iron. Each insert 213 is an annular member. Thereby, each insert 213 has a substantially circular hole 213h. The left case member 211 and the right case member 212 have substantially circular holes 211h and 212h, respectively.

  One insert 213 is cast into the hole 211h of the left case member 211. Thereby, one insert 213 is integrally provided in the left case member 211. The other insert 213 is cast into the hole 212h of the right case member 212. Thereby, the other insert 213 is integrally provided in the right case member 212. As a result, the durability and rigidity of the portion of the crankcase 210 to which the bearings 230 and 240 are attached can be improved.

  The bearing 230 is supported in the hole 213h by one insert 213, and the bearing 240 is supported in the hole 213h by the other insert 213. The outer ring 242 of the bearing 240 is preferably attached to the insert 213 by an interference fit. In this example, the bearing 240 is press-fitted into the hole 213h of the other insert 213. Thereby, the radial bearing gap 240 c of the bearing 240 can be made smaller than the radial bearing gap 230 c of the bearing 230.

  5 is a cross-sectional view taken along line AA in FIG. Hereinafter, the direction of reciprocation of the piston 260 is referred to as a combustion load direction, and is indicated by an arrow in FIG. A direction orthogonal to the combustion load direction is referred to as a load orthogonal direction. The dimension in the combustion load direction of each insert 213 is designed larger than the dimension in the load orthogonal direction. In the example of FIG. 5, the insert 213 has an elliptical outer periphery. The insert 213 is disposed such that the long axis and the short axis are parallel to the combustion load direction and the load orthogonal direction, respectively. The insert 213 may have an oval outer periphery.

  When the outer peripheral portion of the insert has a substantially circular shape, tensile stress acts on the iron insert due to thermal expansion of the aluminum left case member 211 and the right case member 212 during the warm-up operation of the motorcycle 100. Thereby, thermal expansion larger than thermal expansion of a single iron member may occur in the insert. In this case, a gap is generated between the bearings 230 and 240 and the insert, and the influence due to the gap is increased. Effects due to the clearance include contact noise between the bearings 230 and 240 and the insert, vibration due to an impact load in the clearance between the bearings 230 and 240 and the insert, and mechanical loss of the crankshaft 220.

  On the other hand, in this example, the dimension of the outer peripheral part of the insert 213 in the combustion load direction is designed to be larger than the dimension of the outer peripheral part of the insert 213 in the load orthogonal direction. That is, the thickness of the insert 213 in the combustion load direction is larger than the thickness in the load orthogonal direction. In this case, since the insert 213 has high rigidity in the combustion load direction, the degree of deformation of the insert 213 due to the tensile stress of the crankcase 210 during thermal expansion is reduced. This prevents a gap from being generated between the bearings 230 and 240 and the insert 213. As a result, the influence caused by the gap between the bearings 230 and 240 and the insert can be reduced.

  In addition, since the degree of deformation of the insert 213 in the direction perpendicular to the load is small, there is almost no gap between the bearings 230 and 240 and the insert 213 in the direction perpendicular to the load. Therefore, there is almost no influence due to the gap between the bearings 230 and 240 and the insert in the direction perpendicular to the load. Therefore, the thickness of the insert 213 in the direction perpendicular to the load can be minimized. Thereby, while being able to prevent the weight of the insert 213 from increasing, the cost of the insert 213 can be reduced.

  FIG. 6 is an enlarged cross-sectional view of the crank webs 223 and 224 and the periphery thereof. As shown in FIG. 6, the crank web 223 includes a crank arm 223A and a balance weight 223B. The crank web 224 includes a crank arm 224A and a balance weight 224B.

  Of the opposing surfaces of the crank arms 223A and 224A, the end regions closest to the piston 260 (FIG. 2) are referred to as end inner side surfaces 223a and 224a, respectively. End inner side surfaces 223a and 224a are formed in a tapered shape so as to incline outward from the crankshaft 220 side to the piston 260 side. Thereby, the space | interval between the edge part inner surface 223a and the side surface of the large end part 270L of the connecting rod 270 expands gradually from the crankshaft 220 side to the piston 260 side. Further, the distance between the end inner side surface 224a and the side surface of the large end 270L of the connecting rod 270 gradually increases from the crankshaft 220 side to the piston 260 side. In this case, the maximum distance t between the end inner side surfaces 223a and 224a and the side surface of the large end 270L of the connecting rod 270 is, for example, 0.1 mm to 0.5 mm.

  The side surfaces of the balance weights 223B and 224B facing outward are referred to as outer surfaces 223b and 224b, respectively. The outer side surfaces 223b and 224b are formed in a tapered shape so as to inwardly incline from the crankshaft 220 side to the opposite side of the piston 260. Thereby, the space | interval between the outer surface 223b and the side surface of the bearing 230 (FIG. 2) expands gradually from the crankshaft 220 side to the opposite side of the piston 260. Further, the distance between the outer surface 224b and the side surface of the bearing 240 (FIG. 2) gradually increases from the crankshaft 220 side to the opposite side of the piston 260.

  When the combustion speed is increased in order to improve the fuel consumption and output of the engine 200, the crank pin 225 may bend due to an increase in the combustion load from the piston 260. Thereby, as shown by the arrow in FIG. 6, when the crankshaft 220 is deformed, the inner surface 223a of the end of the crank arm 223A approaches the side surface of the large end 270L of the connecting rod 270, and the inner surface of the crank arm 224A 224a approaches the side surface of the large end 270L of the connecting rod 270. Further, the outer surface 223b of the balance weight 223B approaches the bearing 230 (FIG. 2), and the outer surface 224b of the balance weight 224B approaches the bearing 240 (FIG. 2).

  In the present embodiment, the end inner side surfaces 223a and 224a of the crank arms 223A and 224A are formed in a tapered shape, so that the end inner side surfaces 223a and 224a are prevented from colliding with the side surface of the large end portion 270L. Moreover, since the outer surfaces 223b and 224b of the balance weights 223B and 224B are formed in a tapered shape, the outer surfaces 223b and 224b are prevented from colliding with the side surfaces of the bearings 230 and 240.

  Thereby, the hitting sound of the crankshaft 220 can be reduced while avoiding an increase in the weight of the crank webs 223 and 224 and a significant structural change. Further, it is possible to reduce the torque loss of the crankshaft 220 due to friction, and to improve the fuel consumption and output.

(3) Effect When the torque of the engine crankshaft is transmitted through the CVT V-belt, the crankshaft is slightly inclined by receiving a pulling force from the V-belt. When the crankshaft receives an impact load due to the explosion pressure from the piston, the radial movement gap of the crank journal supported by the bearing at a position far from the V-belt is a crank supported by the bearing at a position near the V-belt. It becomes larger than the radial movement gap of the journal. Therefore, the influence of the hitting sound caused by the movement of the crank journal supported by the bearing in the radial direction at a position far from the V-belt is larger than the influence on the hitting sound of other parts.

  According to the configuration of engine 200 according to the present embodiment, radial bearing gap 240 c of bearing 240 is smaller than radial bearing gap 230 c of bearing 230. Therefore, even when the crankshaft 220 receives a pulling force from the V-belt 340, the radial movement gap of the crank journal 222 supported by the bearing 240 is reduced. Thereby, the hitting sound due to the radial displacement of the crankshaft 220 is reduced. In this case, it is not necessary to provide an additional structure in the crankcase 210 in order to reduce the amount of displacement of the crankshaft 220. As a result, it is possible to reduce the hitting sound of the crankshaft 220 while suppressing an increase in the size of the crankcase 210 of the engine, a complicated structure, and an increase in weight.

(4) Other Embodiments (a) In the above embodiment, the outer diameter of the bearing 230 is larger than the outer diameter of the bearing 240, and the rigidity of the bearing 230 is higher than the rigidity of the bearing 240, but is not limited thereto. . The outer diameter of the bearing 240 may be larger than the outer diameter of the bearing 230, and the rigidity of the bearing 240 may be higher than the rigidity of the bearing 230. In this case, the vibration of the crank journal 222 supported by the bearing 240 is reduced. Thereby, the hitting sound of the crankshaft 220 can be reduced.

  (B) In the above-described embodiment, the bearings 230 and 240 are ball bearings that are rolling bearings. However, the present invention is not limited to this, and one or both of the bearings 230 and 240 may be rolling bearings other than the ball bearings. .

  FIG. 7 is a schematic diagram showing another example of the combination of the bearings 230 and 240. In the example of FIG. 7, the bearings 230 and 240 are roller bearings. The radial internal gap of the bearing 240 is smaller than the radial internal gap of the bearing 230. Generally, the rigidity of a roller bearing is higher than that of a ball bearing. Therefore, in the example of FIG. 7, the rigidity of the bearings 230 and 240 can be increased as compared with the example of FIG. Thereby, vibrations of the crank journals 221 and 222 supported by the bearings 230 and 240 are further reduced. Further, the bearings 230 and 240 can be reduced in size and weight. As a result, the crankcase 210 can be reduced in size and weight.

  FIG. 8 is a schematic diagram showing still another example of the combination of the bearings 230 and 240. In the example of FIG. 8, the bearing 230 is a ball bearing and the bearing 240 is a roller bearing. The radial internal gap of the bearing 240 is smaller than the radial internal gap of the bearing 230. As described above, the rigidity of the roller bearing is higher than that of the ball bearing. Therefore, in the example of FIG. 8, the rigidity of the bearing 240 can be made higher than the rigidity of the bearing 230. Thereby, the vibration of the crank journal 222 supported by the bearing 240 is further reduced. Further, the bearing 240 can be reduced in size and weight. As a result, the crankcase 210 can be reduced in size and weight.

  (C) In the said embodiment, although the bearings 230 and 240 are rolling bearings, it is not limited to this. For example, the bearings 230 and 240 may be sliding bearings. The movement gap when using plain bearings as the bearings 230 and 240 is a gap between the outer peripheral surface of the crankshaft 220 and the inner peripheral surface of the bearing. FIG. 9 is a horizontal sectional view of the power transmission device 500 as viewed from above when the bearings 230 and 240 are plain bearings. The configuration of the CVT 300 is the same as the configuration of the CVT 300 in FIG.

  As shown in FIG. 9, when the bearings 230 and 240 are sliding bearings, the engine 200 further includes a pair of housings 280. One housing 280 is held by the left case member 211 by being fitted into the hole 213 h of the insert 213 of the left case member 211. The other housing 280 is held by the right case member 212 by being fitted into the hole 213 h of the insert 213 of the right case member 212.

  The bearings 230 and 240 are held by the housing 280. FIG. 10 is an enlarged cross-sectional view of the bearings 230 and 240 when the bearings 230 and 240 are sliding bearings. In this example, as shown in FIG. 10A, a gap (for example, oil clearance) between the outer peripheral surface of the crank journal 221 of the crankshaft 220 and the inner peripheral surface of the bearing 230 is a radial bearing gap 230c. Become. Similarly, as shown in FIG. 10B, the gap between the outer peripheral surface of the crank journal 222 of the crankshaft 220 and the inner peripheral surface of the bearing 240 becomes a radial bearing gap 240c.

  Sliding bearings have higher rigidity than other types of bearings such as rolling bearings. Therefore, by using a sliding bearing as the bearings 230 and 240, vibrations in the frequency range where the hitting sound of the crankshaft 220 can be reduced. Thereby, the hitting sound of the crankshaft 220 can be reduced with a simple configuration. The bearing 240 is preferably attached to the housing 280 provided in the right case member 212 by an interference fit. Thereby, the play of the bearing 240 in the crankcase 210 is prevented. As a result, the hitting sound of the crankshaft 220 can be reduced.

  (D) FIG. 11 is a schematic diagram showing the relationship between the combustion load direction and the direction of the tensile force by the V-belt 340. In the above embodiment, as shown in FIG. 11A, the combustion load direction and the direction of the tensile force by the V-belt 340 are substantially parallel. The combustion load direction is substantially the same as the direction in which the V belt 340 extends, that is, the direction connecting the driving pulley 320 and the driven pulley 330. In this case, the hitting sound due to the radial displacement of the crankshaft 220 is effectively reduced.

  The relationship between the combustion load direction and the direction of the pulling force by the V-belt 340 is not limited to the example of FIG. In the example of FIG. 11B, the direction of the combustion load and the direction of the tensile force by the V belt 340 intersect at an angle of 90 ° or close thereto. The present invention is also applicable to an engine having such a configuration. Also in this case, it is possible to reduce the hitting sound due to the radial displacement of the crankshaft 220.

  (E) In the above embodiment, the V belt 340 functions as a power transmission member, but the power transmission member is not limited to the V belt 340. FIG. 12 is a schematic diagram showing the configuration of the engine 200 when a gear group is used as a power transmission member.

  In the engine 200 of FIG. 12, a drive gear 400 is attached near one end of the crankshaft 220. A driven gear 410 is attached to the main shaft 120. The drive gear 400 and the driven gear 410 are provided so as to mesh with each other. In this example, the drive gear 400 is a power transmission member. The bearing 240 in FIG. 4 is provided at a position farther from the drive gear 400 than the bearing 230.

  When the crankshaft 220 rotates together with the drive gear 400 in one direction indicated by arrow a, the driven gear 410 rotates in the reverse direction indicated by arrow b. In this case, an upward force acts on the drive gear 400 and the driven gear 410. Thereby, one end of the crankshaft 220 receives a downward reaction force indicated by an arrow Y.

  Even in such a case, in the above configuration, since the radial bearing gap 240c of the bearing 240 is smaller than the radial bearing gap 230c of the bearing 230, the radial movement gap of the right end portion of the crankshaft 220 is reduced. Is done. Thereby, the hitting sound due to the radial displacement of the crankshaft 220 can be reduced.

  (F) In the above embodiment, the bearings 230 and 240 are provided on one side of the V-belt 340 or the drive gear 400 that is a power transmission member. In the present invention, the bearings 230 and 240 are cranked so as to sandwich the power transmission member. The present invention can also be applied to a structure provided on the shaft 220. When the distance between the bearing 240 and the power transmission member is longer than the distance between the bearing 230 and the power transmission member, the bearing gap 240c of the bearing 240 is set smaller than the bearing gap 230c of the bearing 230. Thereby, the hitting sound due to the radial displacement of the crankshaft 220 can be reduced.

  (G) Although the motorcycle 100 has been described as an example of the vehicle in the above embodiment, the present invention is not limited to this. The present invention may be applied to other vehicles such as an automatic tricycle or ATV (All Terrain Vehicle).

(5) Correspondence between each component of claim and each part of embodiment The following describes an example of the correspondence between each component of the claim and each part of the embodiment. It is not limited.

  In the above embodiment, the rear wheel 7 is an example of a wheel, the engine 200 is an example of an engine, the V-belt 340 or the drive gear 400 is an example of a power transmission member, and the crankshaft 220 is an example of a crankshaft. It is. The bearings 230 and 240 are examples of first and second bearings, the bearing gaps 230c and 240c are examples of bearing gaps, and the housing 280 is an example of first and second housings. The crankcase 210 is an example of a crankcase, the CVT 300 is an example of a belt-type continuously variable transmission, the V-belt 340 is an example of a belt, the vehicle body 1 is an example of a main body, and the motorcycle 100 is an example of a vehicle. It is an example.

  As each constituent element in the claims, various other elements having configurations or functions described in the claims can be used.

  The present invention can be effectively used for an engine and a vehicle including a bearing that supports a crankshaft.

DESCRIPTION OF SYMBOLS 1 Car body 2 Front fork 3 Front wheel 4 Handle 5 Seat 6 Control device 7 Rear wheel 100 Motorcycle 110 Starting clutch 120 Main shaft 130 Gear mechanism 140 Rear wheel shaft 200 Engine 210 Crankcase 211 Left case member 211h to 213h, 230h, 240h Hole 212 Right case member 213 Insert 220 Crankshaft 221, 222 Crank journal 223, 224 Crank web 223A, 224A Crank arm 223a, 224a End inner side 223B, 224B Balance weight 223b, 224b Outer side 225 Crank pin 230, 240 Bearing 230c, 240c Bearing Gap 231, 241 Inner ring 231 c, 232 c, 241 c, 242 c Gap 232, 242 Outer ring 233, 243 Rolling member 2 34 Holder 250 Cylinder 260 Piston 270 Connecting rod 270L Large end 280 Housing 300 CVT
310 Transmission Case 320 Drive Pulley 330 Driven Pulley 340 V Belt 400 Drive Gear 410 Driven Gear 500 Power Transmission Device

Claims (11)

An engine that generates torque for rotating wheels,
A crankshaft to which a power transmission member for transmitting torque for driving the wheels is attached;
The same type of first and second bearings supporting the crankshaft,
The second bearing is located farther from the power transmission member than the first bearing,
An engine in which a radial bearing clearance of the second bearing is smaller than a radial bearing clearance of the first bearing. The first and second bearings are first and second rolling bearings, respectively.
The radial bearing gap of the first bearing is a radial internal gap of the first rolling bearing, and the radial bearing gap of the second bearing is a radial internal gap of the second rolling bearing, The engine according to claim 1. The engine according to claim 2, wherein a radial internal clearance of the second bearing is 0 or less in a state where the crankshaft and the second bearing are assembled as an engine. The engine according to claim 2 or 3, wherein an inner ring of the second bearing is attached to the crankshaft by an interference fit. A crankcase for holding the first and second bearings;
The engine according to any one of claims 2 to 4, wherein an outer ring of the second bearing is attached to the crankcase by an interference fit. The first and second bearings are first and second plain bearings, respectively.
The radial bearing gap of the first bearing is a gap between the inner surface of the first sliding bearing and the outer peripheral surface of the crankshaft, and the radial bearing gap of the second bearing is The engine according to claim 1, wherein the engine is a gap between an inner surface of a second plain bearing and an outer peripheral surface of the crankshaft. First and second housings for holding the first and second bearings;
A crankcase for holding the first and second housings,
The engine of claim 6, wherein the second bearing is attached to the second housing by an interference fit. The engine according to any one of claims 1 to 7, wherein the rigidity of the second bearing is higher than the rigidity of the first bearing. The engine according to any one of claims 1 to 7, wherein the rigidity of the first bearing is higher than the rigidity of the second bearing. The engine according to any one of claims 1 to 9, wherein the power transmission member is a belt of a belt type continuously variable transmission. The engine according to any one of claims 1 to 10,
A vehicle comprising: a main body that moves by torque generated by the engine.

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