JP6336115B2 - Balancer device for internal combustion engine - Google Patents

Description

  The present invention relates to a balancer device that cancels secondary vibrations of an internal combustion engine.

  In reciprocating engines (hereinafter simply referred to as engines) mounted in automobiles, etc., two balancer shafts each having a balancer weight (counter weight) are arranged to cancel the secondary vibration generated by the piston, thereby transmitting power. A balancer device is installed to drive one balancer shaft by rotation of the crankshaft transmitted through the mechanism and to drive the other balancer shaft connected to each other by gears at the same speed in the opposite direction. May be.

  As such a balancer device, a casing consisting of an upper casing and a lower casing for receiving a pair of balancer shafts is disposed in an oil pan below the cylinder block, and an oil pump body is integrally formed in the lower casing, A pair of balancer shafts are geared together, and driving force is transmitted from the crankshaft via a sprocket provided at one end of one balancer shaft, an endless chain, and a sprocket provided at one end of the crankshaft. A configuration is known in which the balancer shafts are rotated in opposite directions at twice the number of rotations of the crankshaft (see Patent Document 1).

  As another configuration, the second sprocket around which the first chain is wound together with the first sprocket attached to the crankshaft is fixed to the first shaft, and the third gear fixed to the first shaft is fixed to the first balancer shaft. The fifth gear fixed to the first balancer shaft and the sixth gear fixed to the second balancer shaft are meshed with each other, and the number of teeth of the first and second sprockets is set to be the same. Rotate at the same speed as the crankshaft, set the number of teeth of the third gear to twice the number of teeth of the fourth gear and rotate the first balancer shaft at twice the number of rotations of the crankshaft. By setting the number of teeth of the six gears to the same number, the first and second balancer shafts are rotated in the opposite direction at the same rotation number, and the engine vertical direction is higher than that of the first and second balancer shafts. Downwards in the piston reciprocating motion direction) (structure in which a first axis position) spaced from the crankshaft are also known (see Patent Document 2).

Japanese Patent No. 4072251 Japanese Patent No. 3707140

  However, when the rotational speed of the crankshaft is doubled by a transmission mechanism in which a chain is wound around two sprockets on the drive side and driven side as in the configuration of Patent Document 1, the driven side (balancer shaft side) Since the sprocket has a small diameter and the number of teeth (the number of teeth) is reduced, the chain cordal action (change in the position of the chain due to polygonal movement) increases. If the driven sprocket has a small diameter, the collision speed of the chain with the sprocket (the collision speed in the sprocket radial direction) increases, and the collision energy increases. For this reason, the meshing vibration generating force of the chain with respect to the sprocket is large, and the noise is increased.

  To reduce the diameter of the driven sprocket, the diameter of the drive sprocket can be increased. However, increasing the diameter of the drive sprocket can be achieved by using auxiliary equipment and auxiliary equipment (in general, crankshaft This is not preferable because it reduces the degree of freedom of layout of a pulley or belt provided on the same side as the balancer chain.

  On the other hand, a first shaft that is rotationally driven by a crankshaft is provided separately from a pair of balancer shafts as in the configuration of Patent Document 2, and the rotational speed is doubled when power is transmitted from the first shaft to the balancer shaft. When the speed is increased, the driven sprocket fixed to the first shaft can be prevented from being reduced in diameter, and noise can be reduced. However, since the diameter of the third gear fixed to the first shaft is twice the diameter of the fourth gear fixed to the first balancer shaft, the first shaft is used in order to ensure a distance from the crankshaft. Therefore, the amount of oil immersion (depth) of the third gear is increased, and friction due to oil agitation increases.

  In view of such a background, the present invention provides a balancer device for an internal combustion engine that can reduce meshing excitation force in a winding transmission mechanism and reduce friction caused by oil agitation in a gear transmission mechanism. Let it be an issue.

  In order to solve such a problem, the present invention is provided in an internal combustion engine (1) and rotates a pair of balancer shafts (12) in opposite directions at a rotational speed twice that of a crankshaft (2). A balancer device (10), and an input shaft (13) connected to the crankshaft via a winding-type first transmission mechanism (16), and at least one pair of gears (13e, 12Fe, 51a) meshing with each other , 51b, 12Re) and a pair of gears (12Fd, 12Rd) having the same number of teeth meshing with the first balancer shaft (12F, 12R) connected to the input shaft via the second transmission mechanism (17). A second balancer shaft (12R, 12F) connected to the first balancer shaft via a third transmission mechanism (18) consisting of The speed increasing ratio of the first transmission mechanism and the second transmission mechanism is set so that the rotational speed of the input shaft is faster than the rotational speed of the crankshaft and slower than the rotational speed of the first balancer shaft. The configuration is as follows.

  According to this configuration, it is possible to reduce the meshing excitation force in the winding-type first transmission mechanism, and it is possible to reduce friction caused by oil agitation in the gear-type second transmission mechanism.

  In the above invention, the first transmission mechanism and the second transmission mechanism may be increased so that the rotational speed of the input shaft is 1.1 to 1.6 times the rotational speed of the crankshaft. The speed ratio is preferably set.

  According to this configuration, it is possible to achieve both a reduction in meshing vibration force and a reduction in friction due to oil agitation in a balanced manner.

  In the above invention, the input shaft may be arranged coaxially with one (12R) of the pair of balancer shafts.

  According to this configuration, it is possible to facilitate the processing of the journal bearing that supports the input shaft.

  In the above invention, the input shaft 13 and the one balancer shaft (12R) are arranged so as to face each other with a slight gap therebetween, and both end portions facing each other have a single journal bearing ( 22) It is good to make it the structure which makes the journal (13c, 12Ra) of the same diameter pivotally supported.

  According to this configuration, it is only necessary to form one journal bearing in order to support the two journals, and therefore processing can be facilitated.

  In the above invention, the journals (13c, 12Ra) of the input shaft and the one balancer shaft (12R) have the same length, and the intermediate position in the axial direction of the journal bearing (22). It is preferable that an oil passage (29) communicating with the gap between the input shaft and the one balancer shaft is formed.

  According to this configuration, since it is not necessary to form an oil groove on each of the bearing surfaces of the journal bearing, it is possible to reduce the axial dimension of the journal bearing that is necessary for securing a supporting load.

  As described above, according to the present invention, it is possible to provide a balancer device for an internal combustion engine that can reduce the meshing excitation force in the winding transmission mechanism and the friction caused by oil agitation in the gear transmission mechanism.

Sectional drawing which shows the engine lower part which concerns on embodiment along a back balancer shaft Sectional view along the line II-II in FIG. Explanatory drawing which shows the power transmission path | route of the balancer apparatus shown in FIG. Illustration of chain chordal action (A) Correlation of collision energy reduction amount to speed increase ratio (B) Graph showing correlation of gear oil immersion depth Schematic diagram of a balancer device according to a modified embodiment

  Hereinafter, an embodiment in which a balancer device 10 according to the present invention is applied to an in-line four-cylinder automobile engine (hereinafter simply referred to as an engine 1) will be described in detail with reference to the drawings.

  As shown in FIG. 1, the engine 1 is an in-line four-cylinder engine in which a crankshaft 2 extends in a horizontal direction, and is mounted on an automobile in a posture in which a cylinder axis is inclined rearward. The vertical direction is determined in a state where the engine 1 is mounted on the automobile. However, in the following description, in order to facilitate explanation and understanding, the cylinder axis direction extending substantially vertically and the crankshaft direction perpendicular thereto Are assumed to be up, down, left and right, and directions are indicated by arrows in FIG. The left and right directions are based on the traveling direction of the automobile on which the engine 1 is mounted, and are opposite to the left and right in the drawing.

  The engine 1 forms a cylinder and has an upper block 3 having a skirt portion at the lower portion, and a lower block 5 coupled to the lower portion of the upper block 3 and defining a crank chamber 4 in cooperation with the skirt portion of the upper block 3. The oil pan 6 is coupled to the lower part of the lower block 5 and defines an oil reservoir below the crank chamber 4. The balancer device 10 is coupled to the lower part of the lower block 5 and disposed inside the oil pan 6. ing. Hereinafter, the upper block 3 and the lower block 5 are collectively referred to as a cylinder block 7.

  The crankshaft 2 has four crankpins 2a (hereinafter referred to as the right side (the left side in the drawing) sequentially connected from a piston pin (not shown) of a piston (not shown) slidably provided in the cylinder. A crank pin 2a), five journals 2b (hereinafter referred to as first to fifth journals 2b in order from the right side) provided between the crank pin 2a, and a crank connecting the journal 2b and the crank pin 2a. The arm 2c and the crank arm 2c are provided with a counterweight 2d integrally formed on the side opposite to the crankpin 2a. The first and fourth crankpins 2a are disposed at the same phase position, and the second and third crankpins 2a are disposed at the same phase position that is 180 degrees out of phase with the first and fourth crankpins 2a. .

  The bearing wall for supporting the first and fifth journals 2b of the crankshaft 2 is constituted by the right wall and the left wall of the cylinder block 7, and the bearing walls for supporting the second to fourth journals 2b are provided in the crank chamber 4. It is comprised by the provided partition.

  The right end of the crankshaft 2 extends further to the right from the first journal 2 b and protrudes from the right wall of the cylinder block 7. The protruding portion includes a relatively small-diameter small sprocket 2e for driving a camshaft (not shown) and a relatively large-diameter large sprocket 2f (drive sprocket) for driving the balancer device 10 on the first journal 2b side. It is fixed in this order. A chain case 9 is provided outside the large sprocket 2f so as to penetrate the crankshaft 2. A crank pulley 2g for driving an auxiliary machine of the engine 1 is fixed to the right end of the crankshaft 2 positioned outside the chain case 9.

  The balancer device 10 reduces the secondary vibration of the engine 1 caused by the reciprocating motion of the piston. As shown in FIG. 2, in this embodiment, the balancer device 10 is integrally provided with an oil pump 11 for pumping oil to each part of the engine 1 and each sliding part of the balancer device 10. The balancer device 10 includes a pair of front and rear balancer shafts 12 (front balancer shaft 12F and rear balancer shaft 12R) disposed in parallel with the crankshaft 2, and an input shaft 13 disposed coaxially with the rear balancer shaft 12R. The balancer housing 14 supports and accommodates the two balancer shafts 12F and 12R and the input shaft 13. Both balancer shafts 12F and 12R are disposed at the same height in the cylinder axial direction.

  The balancer housing 14 is coupled to an upper housing 14A and a lower housing 14B which are divided into two vertically along a plane passing through the axial centers of the balancer shafts 12F and 12R, and a right end surface of the lower housing 14B and the upper housing 14A. The right housing 14 </ b> C constituting the pump body 11 a of the pump 11. A pump cover 11b is coupled to the right end surface of the pump body 11a of the right housing 14C. The balancer device 10 is fastened to the lower surface of the lower block 5 (below the crankshaft 2) by a through bolt inserted from below into a bolt insertion hole provided at an appropriate position of the balancer housing 14.

  The input shaft 13 is provided so as to protrude rightward from the balancer housing 14, and a driven sprocket 13a is fixed at a position corresponding to the large sprocket 2f in the crankshaft direction of the protruding portion. The input shaft 13 has a first journal 13b on the left side of the driven sprocket 13a and a second journal 13c on the left end of the shaft extending leftward from the first journal 13b.

  The first journal 13b of the input shaft 13 is supported by a first journal bearing 21 formed so as to penetrate the right housing 14C, and the second journal 13c of the input shaft 13 is formed in the upper housing 14A and the lower housing 14B. Further, it is supported by a second journal bearing 22 constituted by a half bearing. The right housing 14C is assembled to the upper housing 14A and the lower housing 14B so that the first journal bearing 21 and the second journal bearing 22 are arranged coaxially.

  A roller chain 15 is wound around the large sprocket 2 f of the crankshaft 2 and the driven sprocket 13 a of the input shaft 13. That is, the large sprocket 2f, the roller chain 15, and the driven sprocket 13a constitute a winding-type first transmission mechanism 16 that transmits the rotational force of the crankshaft 2 to the input shaft 13. The input shaft 13 rotates in the same direction as the crankshaft 2.

  On the left side of the first journal 13b of the input shaft 13, a bowl-shaped thrust plate 13d (color) is integrally formed. The inner surfaces of the driven sprocket 13a and the thrust plate 13d sandwiching the first journal bearing 21 are thrust surfaces. That is, the bearing wall constituting the first journal bearing 21 also serves as a thrust bearing that supports the axial load of the input shaft 13. A relatively large first helical gear 13e is fixed between the thrust plate 13d of the input shaft 13 and the second journal 13c.

  In addition to the second journal bearing 22, the upper housing 14 </ b> A and the lower housing 14 </ b> B have a third journal bearing 23 configured by a half bearing formed in the upper housing 14 </ b> A and the lower housing 14 </ b> B in the same manner as the second journal bearing 22. A fourth journal bearing 24 and a fifth journal bearing 25 are formed. The third journal bearing 23 is disposed coaxially with the second journal bearing 22 and spaced leftward from the second journal bearing 22 and below the third journal 2 b of the crankshaft 2. The fourth journal bearing 24 and the fifth journal bearing 25 are coaxially arranged in front of these at the same position in the left-right direction as the second journal bearing 22 and the third journal bearing 23, respectively. The bearing walls forming the second journal bearing 22 and the fourth journal bearing 24 and the bearing walls forming the third journal bearing 23 and the fifth journal bearing 25 are each configured as an integral wall continuous in the front-rear direction. The The second to fifth journal bearings 22 to 25 have substantially the same width dimension.

  The upper housing 14A and the lower housing 14B include three bolt holes 14a formed in the bearing walls forming the second and fourth journal bearings 22 and 24, and bearing walls forming the third and fifth journal bearings 23 and 25. The bolts are fastened to each other by six bolts (not shown) inserted into the three bolt holes 14a. The bolt holes 14a are arranged between the two journal bearings and outside the two journal bearings in each bearing wall. The bolt hole 14a of the upper housing 14A is configured as a bolt insertion hole penetrating the upper housing 14A, and the bolt hole 14a of the lower housing 14B is configured as a bottomed female screw hole into which the bolt is screwed. To be inserted into the bolt hole 14a.

  Both balancer shafts 12R, 12F are supported by first journals 12Ra, 12Fa supported by corresponding second journal bearings 22 or fourth journal bearings 24, and third journal bearings 23 or fifth journal bearings 25, respectively. Second journals 12Rb and 12Fb. Further, the balancer shafts 12R and 12F are provided on the left and right sides of the second journals 12Rb and 12Fb, respectively, and a pair of left and right balancer weights 12Rc having substantially the same shape in which the center of gravity is biased radially outward from the rotation center. 12Fc, and right balancer weights 12Rc, 12Fc and first helical gears 12Rd, 12Fd fixed between the first journals 12Ra, 12Fa. A bearing metal 28 is installed on the third journal bearing 23 and the fifth journal bearing 25 that pivotally support the second journals 12Rb and 12Fb between the left and right balancer weights 12Rc and 12Fc.

  The second journal bearing 22 pivotally supports both the second journal 13c of the input shaft 13 and the first journal 12Ra of the rear balancer shaft 12R. The second journal 13c of the input shaft 13 and the first journal 12Ra of the rear balancer shaft 12R have the same diameter and the same length, and face each other with a slight gap at an intermediate position in the length direction of the second journal bearing 22. Are arranged to be. Accordingly, the first journal 12Ra of the rear balancer shaft 12R is about half the length of the second journal 12Rb of the rear balancer shaft 12R and the first journal 12Fa of the front balancer shaft 12F.

  In both the balancer shafts 12R and 12F, the opposing portions of the pair of left and right balancer weights 12Rc and 12Fc have a bowl-like shape whose diameter is larger than that of the second journals 12Rb and 12Fb (see FIG. 1). The opposed inner surfaces of the bowl-shaped portions are thrust surfaces that transmit a thrust force to the third journal bearing 23 or the fifth journal bearing 25. That is, the bearing walls forming the third and fifth journal bearings 23 and 25 also serve as thrust bearings that support the axial loads of the balancer shafts 12R and 12F.

  As shown in FIG. 2, the rear journal shaft 12 </ b> R has the first journal 12 </ b> Ra constituting the right end. On the other hand, the front balancer shaft 12F extends further to the right from the first journal 12Fa, and a second helical gear 12Fe that meshes with the first helical gear 13e of the input shaft 13 is fixed to the extended portion. That is, the first transmission gear 17 that transmits the rotational force of the input shaft 13 to the front balancer shaft 12F is configured by the first helical gear 13e and the second helical gear 12Fe. As a result, the front balancer shaft 12F rotates in the direction opposite to the input shaft 13. Note that the directions of twisting of the second helical gear 12Fe and the first helical gear 12Fd of the front balancer shaft 12F are the same, thereby reducing the axial load of the front balancer shaft 12F.

  The first helical gear 12Fd of the front balancer shaft 12F and the first helical gear 12Rd of the rear balancer shaft 12R are engaged with each other, and the rotational force of the front balancer shaft 12F is applied to the rear balancer shaft 12R by these first helical gears 12Fd, 12Rd. A third transmission mechanism 18 for transmission is configured. The first helical gears 12Fd and 12Rd constituting the third transmission mechanism 18 have the same diameter and the same number of teeth (speed-up gear ratio = 1), and both balancer shafts 12F and 12R are the same in opposite directions. It rotates at the rotation speed.

  On the other hand, the speed increasing ratio of the first transmission mechanism 16 and the second transmission mechanism 17 is set so that the balancer shafts 12F and 12R have a rotational speed twice that of the crankshaft 2. Specifically, in the present embodiment, the chain transmission speed ratio of the first transmission mechanism 16 is set to 4/3, the speed increase gear ratio of the second transmission mechanism 17 is set to 3/2, and the first transmission mechanism The speed increasing ratio of the mechanism obtained by multiplying 16 and the second transmission mechanism 17 is 2.

  Accordingly, the diameter and the number of teeth (the number of teeth) of the driven sprocket 13a are set to 3/4 times the diameter and the number of the large sprocket 2f, and the rotational speed of the input shaft 13 is twice that of the crankshaft 2. It is larger than the ratio (1/2 times). On the other hand, the diameter and the number of teeth of the first helical gear 13e of the input shaft 13 are 3/2 times the diameter and the number of teeth of the second helical gear 12Fe of the front balancer shaft 12F, and the rotational speed of the input shaft 13 is the crankshaft. It is smaller than the ratio (twice) when it becomes the same as the rotational speed of 2.

  The right end of the front balancer shaft 12F extends further to the right from the second helical gear 12Fe and is connected to a pump shaft 11c that drives the oil pump 11 via a joint. The oil pump 11 includes a pump body 11a defining a cylindrical pump chamber by a right housing 14C and a pump cover 11b, an outer rotor 11d and an inner rotor 11e built in the pump chamber, and a pump shaft 11c fixed to the inner rotor 11e. It is a trochoid type which has a publicly known composition provided with. The pump shaft 11c is pivotally supported by a sixth journal bearing 26 formed through the wall of the right housing 14C. The joint is composed of a key groove 11f formed on the left end surface of the pump shaft 11c, and a key 12Ff which is formed to protrude from the right end surface of the front balancer shaft 12F and fits in the key groove 11f. The front balancer shaft 12F can be easily assembled by being fitted into the key groove 11f.

  The oil pump 11 sucks engine oil from an intake port of an oil strainer (not shown) formed on the bottom wall of the balancer housing 14 and supplies the engine oil to each part of the engine 1 and each sliding part of the balancer device 10 through a discharge passage. Pump. Specifically, in the balancer device 10, engine oil is supplied to the first to fifth journal bearings 21 to 25. In the first, third, and fifth journal bearings 21, 23, and 25, the flowing engine oil is supplied to thrust bearings formed on both side surfaces thereof. In the second journal bearing 22, an oil passage 29 communicates with a gap formed between the input shaft 13 and the rear balancer shaft 12R.

  In the balancer device 10 configured as described above, power is transmitted as shown by an arrow in FIG. That is, the rotational force of the crankshaft 2 is transmitted to the input shaft 13 via the large sprocket 2 f (FIG. 1), the roller chain 15 (FIG. 1) and the driven sprocket 13 a of the first transmission mechanism 16, and the second transmission mechanism 17. Is transmitted to the front balancer shaft 12F via the first helical gear 13e and the second helical gear 12Fe. As described above, the front balancer shaft 12F rotates in the opposite direction to the crankshaft 2 at a rotational speed twice that of the crankshaft 2. The rotational force of the front balancer shaft 12F is transmitted to the pump shaft 11c via a joint, and is also transmitted to the rear balancer shaft 12R via both first helical gears 12Fd and 12Rd constituting the third transmission mechanism 18. The pair of front and rear balancer shafts 12F and 12R rotate at the same rotational speed in opposite directions. As a result, an inertial force in the cylinder axis direction that cancels the secondary vibration of the engine 1 is generated.

  Next, the speed increasing ratio of the first transmission mechanism 16 and the second transmission mechanism 17 will be described.

  First, with reference to FIG. 4, the mechanism of noise generation in the winding-type first transmission mechanism 16 that performs chain transmission using a transmission chain (chain) will be described. 4A and 4B are schematic diagrams for explaining the chordal action of the roller chain 15. FIG. 4A is the largest radius of the roller chain 15 (the speed V of the roller chain 15 when the rotational speed of the driven sprocket 13a is constant). (B) shows a state in which the radius of the roller chain 15 is the smallest (when the rotational speed of the driven sprocket 13a is constant, the speed V of the roller chain 15 is the lowest). Here, it is assumed that the driven sprocket 13a has six tooth portions 31 and six valley portions 32, and the entrance side and the exit side of the roller chain 15 are parallel (the wrap angle is 180 degrees). To do. The driven sprocket 13 a rotates at an angular velocity ω, and the roller 15 a of the roller chain 15 is seated on the valley portion 32. The distance from the rotation center O of the driven sprocket 13a to the center of the roller 15a seated in the valley portion 32 is the maximum radius R, and a straight line connecting the centers of the rollers 15a seated in the valley portion 32 from the rotation center O of the driven sprocket 13a. The length of the perpendicular drawn with respect to q is the minimum radius r.

  As shown in FIG. 4A, in the state where the valley 32 of the driven sprocket 13a is located directly above and below the rotation center O (starting point and ending point of winding), the speed V of the roller chain 15 is The maximum value Vmax = Rω. On the other hand, as shown in FIG. 4B, when the toothed portion 31 of the driven sprocket 13a is located directly above and below the rotation center O (starting point and ending point of winding), the speed V of the roller chain 15 Is the minimum value Vmin = rω. That is, as the driven sprocket 13a rotates, the speed V of the roller chain 15 repeatedly fluctuates in the range from the minimum value Vmin = rω to the maximum value Vmax = Rω for each pitch of the tooth portion 31 of the driven sprocket 13a.

  The fluctuation of the speed V of the roller chain 15 accompanying the polygonal movement of the driven sprocket 13a (also referred to as “string displacement”) becomes the vibration generating force. In other words, when the speed V of the roller chain 15 is constant, the fluctuation of the rotational speed generated in the driven sprocket 13a becomes the vibration generating force. This vibration force vibrates the bearing of the driven sprocket 13a, the cover of the transmission mechanism, and the like, which is one of the causes of noise generation. Further, the difference between the maximum radius R and the minimum radius r (R−r) increases the collision between the driven sprocket 13a and the roller chain 15 to generate a striking sound and vibrate the bearing of the driven sprocket 13a. Will occur.

  As for the noise generated by such a principle, the smaller the radius of the driven sprocket 13a, the smaller the number of teeth (the number of teeth), and the speed fluctuation amount (Rω−rω) and the radius difference (R−r) of the roller chain 15 are reduced. It grows because it grows.

  Therefore, in the present invention, the driven sprocket 13a is enlarged in diameter, and the speed increasing ratio of the first transmission mechanism 16 is made smaller than 2, thereby reducing the polygonal behavior that is the cause of the exciting force indicated by the bold line, As a result, the rotational speed of the driven sprocket 13a is reduced to reduce noise generation.

  On the other hand, in order to make the speed increasing ratio of the first and second power transmission mechanisms 16, 17 closer to 1 as the speed increasing ratio of the first power transmission mechanism 16 approaches 1 by increasing the diameter of the driven sprocket 13a, the second power transmission mechanism The speed increasing ratio of 17 must be close to 2, and the diameter of the first helical gear 13e of the input shaft 13 is increased. When the diameter of the first helical gear 13e is increased, the input shaft 13 has to be disposed below in order to ensure the distance between the first helical gear 13e and the crankshaft 2, and the amount of oil immersion (depth) of the first helical gear 13e must be reduced. ) Increases and friction due to oil agitation increases.

  FIG. 5A shows a collision energy reduction amount with respect to the speed increase ratio of the first transmission mechanism 16 (a collision energy reduction amount based on a speed increase ratio of 2 as a reference. The larger the value, the smaller the collision energy itself. FIG. 5B shows the oil immersion depth of the first helical gear 13e of the input shaft 13 with respect to the speed increasing ratio of the first transmission mechanism 16 (the friction increases as the oil immersion depth increases). . As shown in FIG. 5A, as the speed increase ratio of the first transmission mechanism 16 decreases, the collision energy reduction amount increases and the collision energy itself decreases. On the other hand, as shown in FIG. 5B, the gear oil immersion depth increases and the friction increases as the speed increase ratio of the first transmission mechanism 16 decreases.

  In order to balance the reduction amount of the collision energy in the first transmission mechanism 16 and the reduction amount of the friction due to oil agitation in the second transmission mechanism 17 in a balanced manner, the speed increasing ratio of the first transmission mechanism 16 is 1.1 or more and It is preferable to set it to 1.6 or less. In the present embodiment, in order to minimize the increase in friction due to oil agitation, the speed increasing ratio of the first transmission mechanism 16 is 4/3 and the speed increasing ratio of the second transmission mechanism 17 is 3 / 2 was set.

  Further, since the speed increase ratio of the first transmission mechanism 16 is set to 4/3, the increase in gear oil immersion depth shown in FIG. As compared with the winding transmission mechanism in which is set to 1, friction due to oil immersion of the first helical gear 13e is reduced.

  As described above, the balancer device 10 includes the input shaft 13 connected to the crankshaft 2 via the winding-type first transmission mechanism 16, the first helical gear 13e and the second helical gear 12Fe that mesh with each other, and the input shaft 13 And a second transmission mechanism 17 that connects one of the balancer shaft 12 and the second transmission mechanism 17 so that the rotational speed of the input shaft 13 is faster than the rotational speed of the crankshaft 2 and slower than the rotational speed of the balancer shaft 12. By setting the speed increasing ratio of the first transmission mechanism 16 and the second transmission mechanism 17, the meshing excitation force in the first transmission mechanism 16 is reduced and the friction caused by oil agitation in the gear-type second transmission mechanism 17 is reduced. Can be reduced.

  Moreover, in the balancer apparatus 10 of this embodiment, the partial wear of the 2nd transmission mechanism 17 can be prevented. In other words, when the speed increasing gear ratio of the second transmission mechanism 17 is set to 1, the teeth of the second helical gear 12Fe of the front balancer shaft 12F are compared with one tooth of the first helical gear 13e that meshes with the input shaft 13. Each time it rotates, it meshes and always meshes with the same tooth. When the speed increasing gear ratio of the second transmission mechanism 17 is set to 2, the tooth of the second helical gear 12Fe is rotated every two rotations of the input shaft 13 with respect to one tooth of the first helical gear 13e engaged with each other. Will be engaged. As described above, a high meshing frequency with respect to one tooth is not preferable because wear due to a manufacturing error or the like is biased to a specific tooth, but the rotational speed of the input shaft 13 is not cranked as in the present embodiment. By setting the speed increasing ratio of the first transmission mechanism 16 and the second transmission mechanism 17 so as to be faster than the rotational speed of the shaft 2 and slower than the rotational speed of the balancer shaft 12, the meshing frequency for one tooth is set. Can be reduced and uneven wear can be prevented.

  As described above, the speed increasing ratio of the first transmission mechanism 16 and the second transmission mechanism 17 is set so that the rotational speed of the input shaft 13 is 1.1 to 1.6 times the rotational speed of the crankshaft 2. Is set, it is possible to achieve both a reduction in meshing vibration force and a reduction in friction due to oil agitation in a balanced manner.

  Furthermore, since the input shaft 13 is disposed coaxially with one of the pair of balancer shafts 12F and 12R, the processing of the second journal bearing 22 that supports the input shaft 13 is facilitated.

  In the present embodiment, the input shaft 13 and the rear balancer shaft 12R are arranged so as to face each other with a slight gap therebetween, and both ends facing each other are pivotally supported by a single second journal bearing 22. The second journal 13c and the first journal 12Ra having the same diameter are formed. Therefore, it is only necessary to form one second journal bearing 22 to support the two journals 13c and 12Ra, and the processing becomes easy.

  In the present embodiment, the second journal 13c of the input shaft 13 and the first journal 12Ra of the rear balancer shaft 12R have the same length, and the input shaft 13 is positioned at an intermediate position in the axial direction of the second journal bearing 22. And an oil passage 29 communicating with the gap between the rear balancer shaft 12R and the rear balancer shaft 12R. Therefore, it is not necessary to form an oil groove on each of the bearing surfaces of the second journal bearing 22, and the axial dimension of the second journal bearing 22 necessary for securing a supporting load can be reduced.

<< Modified Embodiment >>
Next, a balancer device 10 according to a modified embodiment will be described with reference to FIG. In addition, the same code | symbol is attached | subjected to the element which has the form or function in common with the said embodiment, and the overlapping description is abbreviate | omitted. The front-rear and left-right directions are in accordance with the above embodiment (see FIG. 2).

  In the above embodiment, in the transmission of power from the crankshaft 2 to the pair of balancer shafts 12F and 12R, the gear-type second transmission mechanism 17 comprising a pair of gears meshing with the winding-type first transmission mechanism 16 and However, in the balancer device 10 of the present modified embodiment, the second transmission mechanism 17 includes two speed increasing mechanisms (two pairs of gears), and the three speeds are increased. Is done.

  This will be specifically described below. The balancer device 10 includes an intermediate shaft 51 arranged in parallel with the input shaft 13 in addition to the pair of front and rear balancer shafts 12F and 12R and the input shaft 13. In the present modified embodiment, the intermediate shaft 51 is disposed coaxially with the front balancer shaft 12F. A first helical gear 51a that meshes with the first helical gear 13e of the input shaft 13 is fixed to the intermediate shaft 51, and a second helical gear 51b is fixed at a position spaced leftward from the first helical gear 51a.

  The configuration of the pair of balancer shafts 12F and 12R is also different from the above embodiment. Specifically, in the front balancer shaft 12F, the first journal 12Fa forms the right end, and the second helical gear 12Fe is not provided. In the rear balancer shaft 12R, a second helical gear that meshes with the second helical gear 51b at a position between the first journal 12Ra and the first helical gear 12Rd and corresponding to the second helical gear 51b of the intermediate shaft 51 in the left-right direction. 12Re is fixed.

  The intermediate shaft 51 includes a first journal 51c at the left end, and is disposed so as to face the right end surface of the front balancer shaft 12F with a slight gap. The first journal 51c of the intermediate shaft 51 and the first journal 12Fa of the front balancer shaft 12F are both supported by the fourth journal bearing 24. On the other hand, the second journal 13c of the input shaft 13 and the first journal 12Ra of the rear balancer shaft 12R are both pivotally supported by the second journal bearing 22 as in the above embodiment.

  The first gear mechanism 17A is constituted by the first helical gear 13e of the input shaft 13 and the first helical gear 51a of the intermediate shaft 51, and the second gear mechanism is constituted by the second helical gear 51b of the intermediate shaft 51 and the second helical gear 12Re of the rear balancer shaft 12R. 17B is configured. The second transmission mechanism 17 is configured by the intermediate shaft 51, the first gear mechanism 17A, and the second gear mechanism 17B.

  In the balancer device 10 configured as described above, power is transmitted as indicated by arrows in the drawing. That is, the rotational force of the crankshaft 2 transmitted to the input shaft 13 via the first transmission mechanism 16 including the driven sprocket 13a is transmitted to the intermediate shaft 51 via the first gear mechanism 17A of the second transmission mechanism 17. The The rotational force of the intermediate shaft 51 is transmitted to the pump shaft 11c and is also transmitted to the rear balancer shaft 12R via the second gear mechanism 17B of the second transmission mechanism 17 so that the rear balancer shaft 12R is twice that of the crankshaft 2. In the same direction as the crankshaft 2. The rotational force of the rear balancer shaft 12R is transmitted to the front balancer shaft 12F via the third transmission mechanism 18, and the front balancer shaft 12F is rotated at the same rotational speed in a direction opposite to the rear balancer shaft 12R.

  The chain transmission ratio of the first transmission mechanism 16 is greater than 1 so that the rotational speed of the input shaft 13 is faster than the rotational speed of the crankshaft 2 and slower than the rotational speed of the rear balancer shaft 12R. It is set to be smaller than 2. The speed increasing gear ratio of the first gear mechanism 17A of the second transmission mechanism 17 is such that the rotational speed of the intermediate shaft 51 is faster than the rotational speed of the input shaft 13 and slower than the rotational speed of the rear balancer shaft 12R. It is set larger than 1 and smaller than 2. For example, when the chain speed increasing ratio of the first transmission mechanism 16 is set to 4/3, the speed increasing gear ratio of the first gear mechanism 17A is 4/3, and the speed increasing gear ratio of the second gear mechanism 17B is 9 /. By setting it to 8, the rotational speed of the balancer shafts 12F and 12R becomes twice the rotational speed of the crankshaft 2.

  Although the description of the specific embodiment is finished as described above, the present invention is not limited to the above embodiment and can be widely modified. For example, in the above-described embodiment, the balancer device 10 for an in-vehicle internal combustion engine has been described as an example. However, the present invention can be widely applied to a railway vehicle, a ship, an aircraft, and the like. Moreover, in the said embodiment, although the helical gear is used for the gear of the 2nd transmission mechanism 17 or the 3rd transmission mechanism 18, you may use a spur gear, a helical gear, etc. Moreover, in the said embodiment, although the roller chain 15 is used for the winding type 1st transmission mechanism 16, you may use the chain of other structures, such as a bush chain. In addition, the specific configuration, arrangement, quantity, angle, and the like of each member and part can be changed as appropriate without departing from the spirit of the present invention. On the other hand, all the elements of the balancer device 10 shown in the above embodiment are not necessarily essential, and can be selected as appropriate.

1 Engine 2 Crankshaft 2f Large sprocket (element of first transmission mechanism 16)
10 balancer device 12F front balancer shaft 12Fd first helical gear (element of third transmission mechanism 18)
12Fe second helical gear (element of second transmission mechanism 17)
12R Rear balancer shaft 12Ra First journal 12Rd First helical gear (element of third transmission mechanism 18)
12Re 2nd helical gear (element of 2nd transmission mechanism 17)
13 Input shaft 13a Driven sprocket (element of first transmission mechanism 16)
13c Second journal 13e First helical gear (element of second transmission mechanism 17)
15 Roller chain (element of first transmission mechanism 16)
16 1st transmission mechanism 17 2nd transmission mechanism 17A 1st gear mechanism 17B 2nd gear mechanism 18 3rd transmission mechanism 22 2nd journal bearing 29 Oil path 51 Intermediate shaft (element of 2nd transmission mechanism 17)
51a First helical gear (element of second transmission mechanism 17)
51b Second helical gear (element of second transmission mechanism 17)

Claims (6)

A balancer device that is provided in an internal combustion engine and rotates a pair of balancer shafts in opposite directions at a rotational speed twice that of a crankshaft,
An input shaft connected to the crankshaft via a winding-type first transmission mechanism;
A first balancer shaft coupled to the input shaft via a second transmission mechanism including at least one pair of gears meshing with each other;
A second balancer shaft connected to the first balancer shaft via a third transmission mechanism comprising a pair of gears having the same number of teeth meshing with each other;
The speed increasing ratio of the first transmission mechanism and the second transmission mechanism is set so that the rotational speed of the input shaft is faster than the rotational speed of the crankshaft and slower than the rotational speed of the first balancer shaft. A balancer device for an internal combustion engine , wherein the first transmission mechanism and the second transmission mechanism together form a speed increasing mechanism .   The speed increasing ratio of the first transmission mechanism and the second transmission mechanism is set so that the rotational speed of the input shaft is 1.1 to 1.6 times the rotational speed of the crankshaft. The balancer device for an internal combustion engine according to claim 1.   The balancer device for an internal combustion engine according to claim 1 or 2, wherein the input shaft is arranged coaxially with one of the pair of balancer shafts.   The input shaft and the one balancer shaft are arranged so as to face each other with a slight gap therebetween, and both ends facing each other form a journal having the same diameter that is pivotally supported by a single journal bearing. The balancer device for an internal combustion engine according to claim 3, wherein the balancer device is an internal combustion engine. The both journals of the input shaft and the one balancer shaft have the same length,
5. The internal combustion engine according to claim 4, wherein an oil passage communicating with the gap between the input shaft and the one balancer shaft is formed at an intermediate position in the axial direction of the journal bearing. Balancer equipment. A pump shaft that drives an oil pump is disposed coaxially with the other of the pair of balancer shafts, and the pump shaft rotates through the joint so that the pump shaft rotates at the same rotational speed as the other of the pair of balancer shafts. The balancer device for an internal combustion engine according to claim 3, wherein the balancer device is connected to the other of the pair of balancer shafts.

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