JPS6347938B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION (A) Field of Industrial Application The present invention relates to a balancer mechanism in a reciprocating engine.

(b) Prior art Conventionally, an approximately elliptical gear link has been interposed between the crankshaft and the balancer shaft to eliminate secondary or higher order reciprocating inertia force or reciprocating inertia couple, thereby creating an engine with less vibration. The technology for doing so is known to the public.

For example, as in Japanese Utility Model Application Publication No. 149345/1983.

There is also a known technique for achieving the same objective by interposing a gear train of non-circular gears. For example, as in Japanese Utility Model Application Publication No. 52-77302.

(c) Problems to be solved by the invention Generally, in a reciprocating engine in which the reciprocating motion of a piston is replaced by rotational motion using a crankshaft, the piston
Large vibrations are generated by the reciprocating inertia force or reciprocating inertia couple generated by the movement of reciprocating parts such as connecting rods and crankshafts.

In order to balance this reciprocating inertia force or reciprocating inertia couple, a balancer that moves in the opposite direction to the movement of the reciprocating part is installed on the balancer shaft to balance the primary reciprocating inertia force or reciprocating inertia couple. It was designed to be balanced.

By using a circular eccentric gear to drive the balancer, the present invention intentionally changes the angular velocity and angular acceleration of the balancer shaft to which the balancer is fixed relative to the rotation angle of the crankshaft, and drives the balancer using a conventional non-eccentric circular gear. This system provides a reciprocating engine with extremely low vibration by balancing secondary or higher order reciprocating inertia forces or higher-order component forces of reciprocating inertia couples that could not be compensated for by conventional balancer devices. It is.

In the present invention, in particular, for this purpose, the meshing arrangement of eccentric gears is arranged so that the acceleration is greatest near top dead center, where the reciprocating inertia force or reciprocating inertia couple of the piston-crank mechanism is greatest. By doing this, the rotation fluctuation of the balancer shaft is increased to increase the reciprocating inertia force or reciprocating inertia couple on the balancer side, which balances the higher-order component of the reciprocating inertia force or reciprocating inertia couple of the piston/crank mechanism. This is what generates the next component.

In addition, in conventional technology, the angular velocity and angular acceleration were intentionally changed using approximate elliptical gears and non-circular gears, but these approximate elliptical gears and non-circular gears cannot be used in normal gear processing machines. It could not be mass-produced and had to be handmade by skilled technicians using special processing machines.

The present invention uses a normal circular gear equipped with a large backlash instead of the approximate elliptical gear or non-circular gear, and simply makes the axis of the circular gear eccentric from the center of the circle. This technology makes it possible to intentionally change angular velocity and angular acceleration.

(d) Means for Solving the Problems The objects of the present invention are as described above, and the configuration for achieving the objects will now be explained.

The reciprocating movement of the piston rotates the crankshaft,
In a reciprocating engine that rotates a balancer shaft from the crankshaft via a gear chain, a pair of circular eccentric gears with an axis located eccentrically from the center are installed on the drive gear chain from the crankshaft to the balancer shaft. In order to maximize the rotational angular velocity of the balancer shaft near the top dead center of the piston, The series is arranged in an interlocking manner.

(E) Embodiment The object and structure of the present invention are as described above. Next, the structure of the embodiment shown in the attached drawings will be explained.

In FIGS. 1 and 2, a piston 1 reciprocates inside a cylinder of an internal combustion engine or the like. The piston 1 is connected to the connecting rod 2 via a piston pin 14. The connecting rod 2 is pivotally connected to the crankshaft 3 via a bearing.

The length of the connecting rod 2 is l1, and the crank radius is l2.

A flywheel 13 is fixed to one end of the crankshaft 3 to eliminate rotational fluctuations of the crankshaft 3 and to provide inertial rotational force.

Further, at the other end of the crankshaft 3, an eccentric gear 4, which is a main part of the present invention, is provided. The eccentric gear 4 is always meshed with the eccentric gear 5 of the idle gear shaft 15.

The idle gear shaft 15 is rotated variably by a combination of eccentric gears 4 and 5, and the idle gear 6 meshes with the balancer gear 7 on the balancer shaft 12.
Further, the balancer gear 7 directly meshes with a balancer gear 8 on another balancer shaft 11.
Therefore, balancer gears 7 and 8 rotate in opposite directions.

In addition, the crankshaft 3 and balancer shafts 11 and 12
rotates once in the same amount of time. A balancer 9 is fixed to the balancer gear 7, and a balancer 10 is fixed to the balancer gear 8.

The balancers 9 and 10 are arranged at exactly opposite positions to the piston 1.

Further, as shown in FIG. 1, the rotation center of the eccentric gear 4 on the crankshaft 3 side is O, the rotation center of the eccentric gear 5 on the idle gear shaft 15 side is O', and both eccentric gears 4, 5
Assuming that the meshing point of is O'', the arrangement is such that O-O'' is maximum and O'-O'' is minimum near top dead center TDC, where the inertial force of the piston-crank mechanism is greatest.

That is, the angular velocity and angular acceleration of the balancer shafts 11 and 12 are at their maximum when the piston is near the TDC position.

FIG. 3 is an enlarged view of the eccentric gears 4 and 5, showing the piston 1 at TDC.

With one rotation of the crankshaft 3, the balancers 9 and 10
Since it is necessary to make one rotation of the eccentric gears 4 and 5, it is sufficient to use the same gears with the same number of teeth, but since the axis of the gear is offset by r1 from the regular center O3 of the circular gear, , as shown in FIGS. 4 to 11, if the pitch circles 4a and 5a of the eccentric gears 4 and 5 are arranged so as to touch each other at the top dead center and bottom dead center positions, the pitch circles will change during the rotation up to that point. This results in a situation where other pitch circles are embedded.

This maximum amount of penetration is shown as maximum penetration amount r3 in FIG.

In order to absorb the maximum biting amount r3, in the present invention, a large backlash is formed in the circular gear.

As shown in FIG. 21, the ratio of this maximum penetration amount r3 to the pitch circle diameter r2 is r3/r2.
When the value of is expressed as the penetration rate, the penetration rate is 0.
~0.1, and the eccentricity ε, i.e. r in Fig. 3.
If the value of 1/r2 is between 0 and 0.10, as shown in Figure 3, the amount of penetration is absorbed by increasing the backlash of the eccentric gear, which is constructed by eccentrically centering the axis of the circular gear. It is possible.

In FIG. 3, 5' is a diagram showing the same gear as the balancer side eccentric gear 5 for comparison, r2 is the pitch circle diameter, and r1 is the amount of eccentricity.

From FIG. 4 to FIG. 11, the eccentric gears 4 and 5 are illustrated with pitch circles 4a and 5a to indicate that the amount of bite occurs.

The maximum biting amount r3 occurs in the engaged state shown in FIG.

From Fig. 12 to Fig. 19, the meshing state of the two eccentric gears 4 and 5 shown in Fig. 3 is arranged on a vertical line, and the eccentric gear 4 on the crankshaft side is rotated in steps of 45 degrees. It shows.

When the crankshaft 3 rotates approximately 45 degrees from the TDC state shown in Figures 4 and 12, the balancer shafts 11 and 1
The idle gear shaft 15 connected to 2 rotates more than 45 degrees.

Thereafter, the rotation angle of the idle gear shaft 15 gradually becomes smaller relative to the approximately 45° rotation of the crankshaft 3, and the rotation of the idle gear shaft near the point where the piston 1 reaches the bottom dead center as shown in FIGS. 8 and 16. The angle of rotation of the crankshaft 3 is approximately 45°, but the rotation angle is less than 45°.

From there, the rotation of the crankshaft gradually increases.
The rotation of the idle gear shaft becomes faster and the piston returns to the top dead center state, that is, the state shown in FIGS. 4 and 12.

In the stages from FIG. 11 and FIG. 19 to FIG. 4 and FIG. 12, the idle gear shaft rotates by more than 45 degrees while the crankshaft rotates by about 45 degrees.

In FIG. 20, the inertia force of the piston-crank mechanism without a balancer is indicated by a.

Also shown is the inertial force b remaining after balancing the inertial force a using a conventional balancer device using a circular gear.

That is, in the conventional balancer device, inertia force b still remains.

Furthermore, by using a balancer device in which the idle gear shaft 15 is rotated by the eccentric gear chains 4 and 5 as in the present invention, changes in the angular velocity and angular acceleration of the balancer occur, so that the reciprocating inertia force or the reciprocating inertia couple A higher order component of the force is created which balances the higher order component of the reciprocating inertia force or reciprocating inertia couple of the piston and crank mechanism.

As a result, when the balancer device of the present invention is provided, the residual inertial force becomes as shown in FIG.

That is, about 96% compared to the case without balancer.
In addition, the inertia force can be reduced by approximately 87% compared to a conventional circular gear-driven balancer device.

This reduction in inertial force results in the effect of reducing vibration.

FIG. 22 shows the optimum range of the rod length/crank radius ratio λ with respect to the eccentricity ε.

The eccentricity ε is a value shown by ε=r1/r2 in FIG. On the other hand, the rod length/crank radius ratio is the value λ=l1/l2 in FIG.

And in order to optimize the balance of inertial forces, ε
It is necessary to specify the value of λ with respect to the value of λ, and in the present invention, the eccentricity ε is specified by the following formula.

ε=−1.134×10 ⁻³ λ ³ +1.640×10 ⁻² λ ² − 9.092×10 ⁻² λ+2.343×10 ⁻¹ .

By setting the eccentricity ε approximately to this value,
This allows for optimal vibration cancellation and smooth rotation.

Let's calculate it as an example. When λ=2, the eccentricity is approximately ε=0.109, and when λ=3, the eccentricity is approximately ε=0.079, λ
When =4, ε=0.06.

In FIGS. 22 and 24, by using eccentric gears 4 and 5 in the drive system of the balancer,
A new eccentric mass will be generated, and the inertia generated by this may cause new vibrations, so even if the eccentric gears 4 and 5 are eccentric, the center of gravity should be statically balanced so that it is not eccentric. Eccentric gears 4, 5 are shown. In particular, in order to reduce the weight in the longitudinal direction, there is a gap 5 in the thick part of the eccentric gear.
b and 5c are provided.

Also, spokes 5d are left so that the strength of eccentric gear 5 is not weakened by gaps 5b and 5c.

On the other hand, on the shorter side, it is not possible to maintain balance just by providing a gap, so protrusions 5a, 5a are provided,
increasing weight. 5e is a hole into which the shaft is inserted. The protrusions may be gradually wedge-shaped or stepped. A weight may be attached instead of the protrusion.

(F) Effect of the Invention The present invention is configured as described above, and the crankshaft 3 rotates as the piston 1 reciprocates between the top dead center and the bottom dead center, but in the conventional case, the balancer shaft rotates. The balancer and the balancer rotated at the same angular velocity as the crankshaft rotated.

Therefore, the primary component of the reciprocating inertia force or the reciprocating inertia couple of the piston-crank mechanism was able to be balanced. However, for other high-order components, it was not possible to balance them if the crankshaft and balancer shaft were rotating at the same time.

In the present invention, since the balancers 8 and 9 rotate at a high angular velocity near the top dead center where the inertial force of the piston/crank mechanism is the largest, the balancers generate a reciprocating inertia force of a higher order component or a reciprocating inertia couple. occurs. This balancer's high-order reciprocating inertia force or reciprocating inertia couple is balanced with the high-order reciprocating inertia force or reciprocating inertia couple of the piston/crank mechanism.

The angular velocity is reduced near the bottom dead center where only a smaller reciprocating inertia force or reciprocating inertia couple is generated.

(G) Effects of the Invention Since the present invention is configured as described above, it has the following effects.

First, in conventional reciprocating engines that are designed to make the fluctuating rotation of the crankshaft a constant rotation on the balancer shaft, there is a high-order component of reciprocating inertia or reciprocating inertia that cannot be eliminated. It was possible to generate a higher-order component on the balancer side to cancel the couple.

Second, in the conventional Japanese Utility Model Application Publication No. 57-149345, an approximate elliptical gear chain is inserted in order to obtain intentional fluctuations in angular velocity and angular acceleration.
Moreover, in Japanese Utility Model Application Publication No. 52-77302, a non-circular gear is used to obtain intentional fluctuations in angular velocity and angular acceleration.

Therefore, when manufacturing approximate elliptical gears and non-circular gears, it is not possible to use the same machine tools that process regular circular gears, and they must be manufactured by hand by skilled workers using special machine tools, which reduces the effectiveness of mass production. Therefore, the cost becomes abnormally high.

On the other hand, in the case of the present invention, the backlash is constructed to be particularly large so as to be able to absorb the maximum biting amount r3 without using an approximate elliptical gear or a non-circular gear. By using a circular gear configured to have a reciprocating width within the range that can be machined, only the shaft center position is made eccentric, making it possible to mass produce at low cost. It is now possible to generate a high-order reciprocating inertia force or reciprocating inertia couple on the balancer 9, 10 side to cancel out the dynamic inertia force or reciprocating inertia couple.

[Brief explanation of the drawing]

FIG. 1 is an overall schematic diagram of a reciprocating engine showing an embodiment of the present invention, FIG. 2 is a side view thereof, FIG. 3 is a drawing showing the meshing state of eccentric gears 4 and 5, and FIGS. Figure 5, Figure 6, Figure 7, Figure 8, Figure 9, Figure 10
Figure 11 shows the state in which the eccentric gear 4 on the crankshaft side rotates approximately 45 degrees at a time, using pitch circles 4a and 5a for gears with the same diameter.
Figure 2, Figure 13, Figure 14, Figure 15, Figure 16
Figures 17, 18, and 19 are drawings showing the rotational state of an embodiment in which the eccentric gear 4 on the crankshaft side is rotatable by increasing its backlash by continuous rotation of about 45 degrees. , FIG. 20 shows the inertial force a without a balancer device, the residual inertial force b after balancing with a conventional balancer device, and the residual inertial force c after balancing with the balancer device of the present invention, respectively. Figure 21 is a diagram showing the relationship between the eccentricity ε and the amount of gear bite, and the shaded area is a diagram showing the range in which ultra-back crush gears can be used, and Figure 22 is a diagram showing the relationship between the eccentricity ε and the amount of gear bite. A drawing showing the optimum range of the eccentricity ε of the eccentric gear with respect to the crank radius ratio (λ), Fig. 23.
FIG. 24 is a plan view and a side cross-sectional view of the eccentric gear in which a gap and a weight are provided to maintain static balance of the eccentric gear itself. 1... Piston, 2... Connecting rod, 3... Crankshaft, 4... Crankshaft side eccentric gear, 5... Balancer side eccentric gear, 9, 10...
... Balancer, 11, 12... Balancer axis.

Claims (1)

[Claims] 1. In a reciprocating engine in which a crankshaft is rotated by the reciprocating motion of a piston, and a balancer shaft is rotated from the crankshaft via a gear train, there is a position eccentric from the center of the drive gear chain from the crankshaft to the balancer shaft. A pair of circular eccentric gears are intermeshed with each other, the shaft center of which is provided at A balancer device for a reciprocating engine, characterized in that an eccentric gear chain is arranged in mesh to maximize the rotational angular velocity of the balancer shaft.