JP3079645B2 - Gear driven balancer for internal combustion engine - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a gear meshing noise reduction structure for a gear driven balancer in a four-cylinder internal combustion engine.

[0002]

2. Description of the Related Art In a four-cylinder internal combustion engine, when the pistons of the first and fourth cylinders are at the top dead center position, the pistons of the second and third cylinders are at the bottom dead center position. An upward unbalance force remains due to the unbalance of the inertial force due to the inertial mass, and a similar upward unbalance force remains even when the pistons of the second and third cylinders are at the top dead center position. Two vertical vibrations occur per rotation of the shaft.

[0003] In order to take the vibration due to the unbalance of the inertial force due to the reciprocating inertial mass of the piston / conrod and the like peculiar to the four-cylinder, it has an unbalance mass that balances the imbalance of the inertial force due to the reciprocal inertial mass of the piston / conrod and the like. A balance shaft is provided and the crankshaft 1
In some cases, a balancer that rotates two times per rotation is provided. However, if only one balance shaft is provided, the vertical vibration caused by the unbalance of the inertial force due to the reciprocating inertial mass of the piston / connecting rod or the like disappears, but the lateral vibration occurs due to the unbalance mass of the balance shaft itself. In order to solve this, two balance shafts are provided, and they are arranged side by side and rotated in reverse to cancel the imbalance in the left-right direction.

[0004] When two balance shafts are arranged below the crankshaft and the balance shaft is rotated in conjunction with the crankshaft, the rotation transmission method is as follows.
Conventionally, there are a chain drive type, a belt drive type, and a gear drive type. For example, Japanese Utility Model Publication No. 58-35880, Japanese Patent Application Laid-Open No. 58-160647, and Japanese Utility Model Laid-Open Publication No. 1-125545 show a gear drive system. The gear drive type has less mechanical loss compared to the chain drive type and the belt drive type, and the balancer can be added relatively easily without adding large fluctuations to the cylinder block structure of the existing balancerless engine. Despite the benefits, there are the following problems that need to be addressed.

[0005]

That is, the gear-driven balancer has a relatively loud noise (hereinafter, referred to as rattle).
(Called a rattle). The rattle noise increases as the distance between the shaft centers between the crank gear and the balancer gear increases and the substantial inter-gear backlash increases.
This is especially noticeable at idle.

There is a limit to reducing backlash between gears in order to reduce rattle noise. For example,
The bearing that supports the balance shaft has a metal clearance that allows the balance shaft to rotate, and because the balance shaft has an unbalanced mass, the balance shaft axis moves circularly within the range of the metal clearance when the balance shaft rotates. The distance between the gear shaft centers fluctuates.
Further, there is a manufacturing tolerance for machining gear teeth. Therefore, when assembling the balancer assembly to the engine, even if the inter-gear backlash is set to be small at one rotational position of the crank gear and the balancer gear, it is set. When the gear rotates, the backlash increases or decreases, and the rattle does not decrease effectively.

In addition, if the distance between the axes of the crank gear and the balancer gear is too small, the backlash becomes zero and the gear may be damaged. For this reason, conventionally, in order to prevent the backlash from being zero, the distance between the gears is set large in consideration of the metal clearance, manufacturing error, and the like, and the distance between the crank gear and the balancer gear for each engine is adjusted. , The balancer was attached to the engine cylinder block. Therefore, noise control was sacrificed.

An object of the present invention is to provide a gear-driven balancer for an internal combustion engine in which, despite the existence of metal clearances and machining errors, these are used in reverse to reduce gear meshing noise. It is in.

[0009]

In order to achieve the above object, in a gear driven crank of an internal combustion engine according to the present invention, rotation of a crankshaft is controlled by a crankshaft mounted on a crankshaft and a balancer gear mounted on a balance shaft. Is transmitted to the balance shaft, and at least one of the crank gear and the balancer gear is attached to the shaft in the following direction.
Before the pitch circle due to a manufacturing error of the at least one gear
The maximum part of the runout amount with respect to the center of the shaft is set so as to be at the meshing position of the crank gear and the balancer gear at the time of the maximum deceleration of the crankshaft.

[0010]

When the generation of rattle noise is analyzed, it is found that rattle noise occurs twice per rotation of the crankshaft, and occurs at the time of the maximum deceleration of the crankshaft (at the time when the speed changes from deceleration to acceleration). In the balancer of the present invention made as described above, at the time of rattle sound generation, the center distance between the crank gear and the balancer gear increases due to the shaft center fluctuation of the balance shaft due to the metal clearance, but at that time, at least at that time, Since the largest part of the tooth groove runout of one gear (the part where the amount of runout of the pitch circle with respect to the rotating shaft center is the largest) comes to the gear meshing position, increase in backlash between gears is suppressed and rattle noise can be reduced. .

Therefore, in the present invention, fluctuations in the distance between the gear shaft centers due to fluctuations in the gear shaft center due to metal clearance and runout of the gear tooth space due to manufacturing tolerances are accepted as being inevitable. By using them in reverse, when the distance between the gears is to be expanded by the metal clearance, the expansion is filled with the runout of the tooth space, and the backlash between the gears is kept small. By selecting the backlash control point as the rattle sound generation point, the most effective rattle sound suppression is achieved.

[0012]

1 to 4 show a preferred embodiment of the present invention. As shown in FIGS. 3 and 4, in a four-cylinder engine, a first balance shaft housing 6 and a second balance shaft housing 8 are fixed to a cylinder block 2 via shims 4. Cylinder block 2
, The crankshaft 10 is rotatably supported, and the first
Between the balance shaft housing 6 and the second balance shaft housing 8 are two balance shafts 1.
2, 14 are rotatably supported.

The two balance shafts 12, 14 are symmetrically disposed below the crankshaft 10. The crankshaft 10 extends in the longitudinal direction of the engine, and the two balance shafts 12, 14 also extend in the longitudinal direction of the engine. The two balance shafts 12 and 14 are disposed near the center of the crankshaft 10 in the longitudinal direction. Each of the two balance shafts 12 and 14 has an unbalance mass 1
6 and 18, the sum of these unbalance masses 16 and 18 balances the imbalance of the inertial force due to the reciprocating inertial mass of the piston / connecting rod and the like in the vertical direction of the engine.

One of the two balance shafts 12, 14 is gear-driven from the crankshaft 10, and the other balance shaft 14 is gear-driven from the one balance shaft 12. The two balance shafts 12, 14 are rotated in opposite directions, thereby balancing the unbalance masses 16, 18 in the left-right direction.

The crankshaft 10 includes a crank gear 2
0 is press-fitted, a balancer gear 22 is press-fitted into the balance shaft 12, and a balancer gear 24 is press-fitted into the balance shaft 14. Then, the crank gear 20 and the balancer gear 22 mesh with each other, and the balancer gear 22 and the balancer gear 24 mesh with each other. The meshing portion between the crank gear 20 and the balancer gear 22 is offset in the engine longitudinal direction from the meshing portion between the balancer gear 22 and the balancer gear 24. The balancer gears 22, 24 are housed in the balance shaft housings 6, 8 so that the rotation of the balancer gears 22, 24 does not stir the oil in the engine oil pan 26.

The gear noise, ie, rattling noise, generated at the meshing portion between the crank gear 20 and the balancer gear 22 depends on the distance between the axis of the balancer gear 22 and the axis of the crank gear 20, and accordingly, the gear noise. It changes according to the substantial inter-gear backlash amount of the crank gear 20. The distance between the gear shaft centers is adjusted by the thickness of the shim 4, and the adjusted distance between the gear shaft centers is determined by the circular motion of the gear shaft within the metal clearance range and the tooth gap of gear teeth due to gear machining errors and the like. It fluctuates when the gear rotates due to runout. In order to reduce the rattle noise, it is, of course, necessary to select the thickness of the shim 4 to be the optimum thickness for suppressing the rattle noise and for preventing gear breakage. It needs to be managed. Then, in order to enable the management of the backlash amount for the reduction of the rattle sound, the generation mechanism of the balancer rattle sound was analyzed.

FIG. 2 shows the result of analyzing the rattle noise generated by the balancer when the engine is idling. As shown in FIG. 2, the rattle is 2 per rotation of the crankshaft 10.
And occurs at a position where the crankshaft 10 rotates 30 ° and 210 ° from the top dead center position of the first cylinder. In terms of crank rotation fluctuation, the positions at 30 ° and 210 ° are the crank's maximum deceleration points, and correspond to the points where the transition from deceleration to acceleration occurs. That is, when the teeth of the crank gear 20 hit the rotation direction front surfaces of the teeth of the balancer gear 22 during deceleration,
Instead of accelerating, the teeth of the crank gear 20 come into contact with the rear side of the teeth of the balancer gear 22 in the rotation direction. However, since there is a backlash gap between the gears, it is considered that a collision occurs and a sound is generated. At the positions of 30 ° and 210 ° of the crankshaft 10, the unbalance mass 16 of the balance shaft 12 is on the far side from the crankshaft 10 as shown in FIG. 2 (however, it is shifted by about 40 ° from the farthest position). Therefore, the axis of the balance shaft 12 is located at a position far from the normal position by the metal clearance from the crankshaft 10, and the backlash between gears is large. For this reason, the collision sound increases. From the above rattle sound analysis, it is the rattle sound generation point,
Two points of the crankshaft maximum deceleration point (30 °, 210 °)
Is considered to be a desirable reference point in reducing noise and managing backlash.

Since the balance shaft 12 has the unbalance mass 16, the axis of the balance shaft 12 makes a circular motion within the range of the metal clearance of the balance shaft bearing. More precisely, the balance shaft 12 is
The centrifugal force by the unbalance mass 16 and the gears 20 and 24
Is pressed against the bearing metal in the direction of the meshing reaction force and the resultant force, and rotates and slides, so that the shaft axis moves circularly. The crankshaft 10 is also unbalanced due to the reciprocating inertial mass of the piston / connecting rod, etc.
The 0 axis makes a circular motion within the metal clearance of the crankshaft bearing. Since the inertial force of the unbalance mass 16 of the balance shaft 12 and the inertial force of the unbalance mass 16 of the balance shaft 12 are balanced with each other in the vertical direction due to the imbalance due to the reciprocating inertial mass of the piston / connecting rod or the like,
When the unbalanced mass force acting on the crankshaft 10 goes upward, the force by the unbalanced mass 16 of the balance shaft 12 goes downward, and the distance between the axis of the crankshaft 10 and the axis of the balance shaft 12 increases. Try to. And 3 which is the rattle sound generation timing
At the positions of 0 ° and 210 °, the unbalance mass 16 of the balance shaft 12 is on the side far from the crankshaft 10 (however, it is shifted by about 40 ° from the farthest point), and the distance between the shafts 10 and 12 is It is time to spread as shown in FIG.

On the other hand, the crank gear 20, the balancer gear 2
2 always has a manufacturing tolerance (manufacturing error less than the tolerance).
There is a runout in the tooth space. Here, the runout of the tooth space is
It is defined as the difference between the maximum value and the minimum value of the radial position reading when the contact pieces such as balls and pins contact the tooth surfaces on both sides of the tooth space near the center of the effective tooth height and when all the tooth spaces are in contact. And substantially corresponds to the amount of radial runout of the gear pitch circle with respect to the rotation axis (the center of the center hole of the gear).

At the stage when the gears 20 and 22 have been manufactured, and at the stage where the shaft has not yet been pressed into the center hole of the gear, a gear meshing tester (the gear to be tested is meshed with a fixed master gear by meshing the gear to be tested). The amount of advance and retreat with respect to the master gear is measured, and it is performed on a plurality of teeth of the gear to determine the runout of the tooth space.
A marking is provided at the circumferential position where the maximum tooth space runout is indicated.

Then, positions corresponding to 30 ° and 210 ° after the top dead center of the crankshaft, which is the rattle sound generation timing, that is, the piston /
30 ° or 21 ° after connecting rod imbalance
When the angle position reached 0 ° and the unbalance mass 16 in terms of the balancer gear 22 came to a position shifted from the farthest point shown in FIG. The relative angular position between the gear and the shaft is determined so that the maximum position of the tooth groove runout is at the meshing point of the crank gear 20 and the balancer gear 22, and the respective gears 20, 22 are maintained so as to maintain this position.
The shafts 10 and 12 are press-fitted into the shafts to be integrated. However, the gear and the shaft may be attached by a method other than press fitting, for example, shrink fit, fitting and key.

Thus, the mounting direction of the crank gear 20 and the balancer gear 22 to the shafts 10 and 12 is
The maximum runout of the tooth grooves 0 and 22 is set so as to be at the meshing position between the crank gear 20 and the balancer gear 22 at the time of the maximum deceleration of the crankshaft. In this case, it is desirable that this setting be performed for both the crank gear 20 and the balancer gear 22. However, in some cases, this setting may be performed for only one of the crank gear 20 and the balancer gear 22, and such a setting may be performed. In any case, the present invention includes the present invention.

As shown in FIG. 1, at the 30 ° and 210 ° positions after the top dead center of the crankshaft, which is the timing at which the rattle sound is generated, the distance between the shaft centers of the gears 20 and 22 is increased (the amount of expansion is d). ₀ ) at this time, the gears 20, 2
Since the maximum runout of the tooth groove 2 comes to the gear meshing position and the runouts d ₁ and d _{2 of the} respective tooth grooves fill in the spread d ₀ , the actual spread amount is zero or negligible. For this reason, the backlash between the gears does not increase, and the rattle noise does not increase or is suppressed to a very small increase even if there is. Thus, the object of the rattle suppression of the present invention is achieved.

As another operation of the present invention, consider the timing at which the unbalance masses 16 and 18 of the balance shafts 12 and 14 approach each other. At this time, contrary to the timing of the rattle sound generation timing, the axes of the balance shafts 12 and 14 approach each other, and the backlash between the gears 22 and 24 becomes extremely small or zero. Absent. However, when the shafts 12 and 14 are press-fitted and used by determining the maximum position of the tooth groove runout as in the present invention, the backlash between the gears 22 and 24 inevitably expands at the time of rattling sound generation,
24 damage can be avoided.

[0025]

According to the present invention, the shafts 10, 1 of at least one of the crank gear 20 and the balancer gear 22 are provided.
2 depends on the manufacturing error of gears 20 and 22
The pitch circle is set so that the largest part of the runout amount with respect to the center of the shaft is at the meshing position of the gears 20 and 22 at the time of the maximum deceleration of the crankshaft (that is, at the time of rattle sound generation timing). Even if the space between the shafts 10 and 12 is widened due to the shaft axis eccentricity caused by the mass, the runout of the tooth space (the pitch circle due to a manufacturing error) may occur.
(Running to the center of the shaft) fills the spread between gears,
Substantial increase in backlash is suppressed, and rattle is reduced.

[Brief description of the drawings]

FIG. 1 is a schematic front view of a gear-driven balancer for an internal combustion engine according to the present invention, showing the spread of inter-gear spread due to runout of tooth spaces.

FIG. 2 is a diagram showing the relationship between crank rotation fluctuation, the rotational position of an unbalance mass of a balance shaft, and the generation timing of rattle sounds, showing the analysis state of rattle sounds.

FIG. 3 is a front view of a gear-driven balancer for an internal combustion engine according to one embodiment of the present invention.

FIG. 4 is a side view of the device of FIG. 3;

[Explanation of symbols]

 2 Cylinder block 4 Shim 6 (First) balance shaft housing 8 (Second) balance shaft housing 10 Crank shaft 12 Balance shaft 14 Balance shaft 16 Unbalance mass 18 Unbalance mass 20 Crank gear 22 Balancer gear 24 Balancer gear

──────────────────────────────────────────────────続 き Continuation of the front page (56) References JP-A-59-133845 (JP, A) JP-A-4-258551 (JP, A) JP-A-3-17357 (JP, U) (58) Survey Field (Int.Cl. ⁷ , DB name) F16F 15/26 F16H 55/17 F02B 77/00

Claims (1)

(57) [Claims] 1. A gear-driven balancer for an internal combustion engine, wherein rotation of a crankshaft is transmitted to a balance shaft by meshing between a crank gear mounted on the crankshaft and a balancer gear mounted on the balance shaft. and wherein at least one of the balancer gear, the mounting direction of the shaft, the shaft of the pitch circle by manufacturing error of the at least one gear
A gear-driven balancer for an internal combustion engine, wherein a maximum portion of a runout amount with respect to a center is set so as to be at a meshing position between the crank gear and the balancer gear when the crankshaft is most decelerated.