JPH0641777B2 - Engine balance mechanism - Google Patents

Description

TECHNICAL FIELD The present invention relates to a balance mechanism for an engine having a uniaxial primary balancer.

(Prior Art) In a conventional engine having a uniaxial primary balancer, the centerline C0 of the cylinder 1 intersects with the rotation center O1 of the crankshaft 3 as shown in FIG. 2 is a 1-axis primary balancer, which is arranged parallel to the crankshaft 3 and
It is linked to the crankshaft 3 through a pair of gears and rotates at the same speed as the crankshaft 3 but in the opposite direction.

By such a method, of the vibration component in the X-axis direction generated by the reciprocating motion of the piston or the like, both the remaining portion of the counterbalancer 7 around the crankshaft 3 that cancels out and the vibration in the Y-axis direction generated by the counter balancer 7 are generated. It is possible to cancel and almost eliminate the primary component of the vibration in the X and Y axis directions.

However, when the uniaxial primary balancer 2 is provided as shown in FIG. 7, it is possible to cancel the above-mentioned primary components of vibration in the X and Y axis directions. Since it acts at a distance of O1 to O2 from the crankshaft center O1, the magnitude of the centrifugal force generated by the balancer 2 about the crankshaft center O1 in the direction perpendicular to the straight line O1O2 and the distances O1 to O2. A new problem arises in that the rotational moment is generated and rotational vibration occurs in a plane perpendicular to the crankshaft 3. In addition, in order to eliminate this rotational vibration, the primary balancer has two axes,
Some of them are arranged on both sides of the output shaft, but they have the problems of high cost and lack of compactness.

For such a problem, for example, by making the cylinder center line eccentric from the crankshaft rotation center toward the primary balancer side, the rotational vibration can be suppressed, but this time the vibration in the cylinder center line direction increases. Remains.

(Object of the Invention) An object of the present invention is to eccentric the cylinder center line to the primary balancer side so that the moment around the crankshaft caused by the uniaxial primary balancer causes the moment around the crankshaft caused by the reciprocating motion of the piston or the like. By canceling out with, reducing the rotational vibration in the plane perpendicular to the crankshaft, and devising the position of the center of gravity of the balancer,
This is to prevent an increase in vibration in the cylinder centerline direction.

(Structure of the Invention) A 1-axis primary balancer that rotates at the same speed as the crankshaft and in the opposite direction is provided parallel to the crankshaft, and the cylinder center line is eccentric to the primary balancer side with respect to the rotation center of the crankshaft. , The center of gravity of the counter balancer provided on the crank web is shifted backward from the position of the phase difference of 180 ° with respect to the crankpin in the rotational direction, and the center of gravity of the primary balancer is described above with respect to the phase difference of 180 ° with respect to the crankpin. It is characterized in that the position is shifted to the rear side in the rotational direction of the primary balancer from the position of symmetrical phase.

(Example) In FIG. 1 showing a balance mechanism of an engine to which the present invention is applied, 10 is a piston, 11 is a cylinder, 12 is a crankshaft, and 13 is a connecting rod. The large end of the control rod 13 is fitted to the crank pin 15 and
Is connected to the crankshaft 12 via a crank web 16. A counter balancer 17 is formed on the crank web 16.

The uniaxial primary balancer 20 is arranged in parallel with the crankshaft 12, is interlockingly connected to the crankshaft by a pair of gears (not shown), and rotates at the same speed as the crankshaft 12 but in the opposite direction. That is, the crankshaft 12 rotates in the direction of arrow A1 and the primary balancer 20 rotates in the direction of arrow A2. Further, the rotation center O2 of the primary balancer 20 and the crankshaft 12
The line connecting the rotation centers O1 of is perpendicular to the center line C0 of the cylinder 11.

The cylinder center line C0 is eccentric to the primary balancer 20 side with respect to the rotation center O1 of the crankshaft 12 by an eccentric amount E.

In FIG. 2, the position of the counter balancer center of gravity C1 is shifted from the position B1 having a phase difference of 180 ° with respect to the clamp pin 15 to the rear side in the rotation direction A1 by an appropriate amount dθ, and 1
The position of the center of gravity C2 of the next balancer is changed to the position B1 having a phase difference of 180 ° with respect to the crank pin 15 and the position B2 having a symmetrical phase.
Is shifted to the rear side in the rotation direction A2 by an appropriate amount dθt.

(Operation) First, the operation of reducing the rotational excitation force around the crankshaft will be described. In FIG. 2, the reciprocating motion of the piston 10 causes the crankshaft 12 to rotate at an angular velocity W in the arrow A1 direction, and the primary balancer 20 to rotate at an angular velocity W in the arrow A2 direction. Due to the movement of each mass point inside the engine, the entire engine causes rotational vibration about a center of gravity in a plane perpendicular to the crankshaft. Since the center of gravity of the engine moves with time and the analysis becomes complicated, the rotational excitation force around the crankshaft is referred to as Mz, and the rotational vibration will be described. Although the piston side pressure and the like actually affect the rotational vibration, they are not included in Mz described below.

The inertial moment Mza about the crankshaft is generated by the rotation of the primary balancer 20, but the cylinder center line C0 is 1
Since it is eccentric to the next balancer 20 side, the piston 10
The moment M about the crankshaft due to the reciprocating weight WREC of the above is generated so as to cancel out the moment Mza, and thereby the rotational excitation force Mz about the crankshaft is greatly reduced. Note that WROT in FIG. 2 is a rotational weight, Wco is a counterweight of the counter balancer 17, and WBI is a counterweight of the primary balancer 20.

When the cylinder center line C0 is not eccentric from the crankshaft rotation center O1 as in the conventional FIG. 6, the rotational excitation force around the crankshaft is equal to the inertial moment Mza about the crankshaft by the primary balancer 20 in FIG. Become. The rotational excitation force Mza of the above conventional example and the rotational excitation force Mzb around the crankshaft when the cylinder center line C0 is eccentric are expressed by the following equations.

^{Mza = WBI × RW 2 cosθ /} g × (Y1 + RBIsinθ) -WBI × RW 2 sinθ / g × (-X1 + RBIcosθ) Mzb = Mza-M = Mza-WREC × RW 2 [cosθ + {λcos2θ + sinθ (λsinθ-ε) 3 + εsinθ} / {1- (λ sin θ −ε) ² } ^3/2 ] g × E WREC ... reciprocating weight, WROT ... rotating weight, WBI ... 1
Next balancer balance weight, R ... crank radius, RBI ... 1
Next balancer radius of gravity, L ... connecting rod length, λ ... continuous rod ratio (R / L), ε ... eccentricity ratio (E / L), θ ... crank angle,
δ ... overbalance rate, g ... gravitational acceleration, X1, Y1
The X and Y coordinates of the primary balancer when the center of rotation of the crankshaft is used as the origin. When the above two expressions are compared, | Mza | max> | Mzb | max.

Normally, the balance weight WB1 of the primary balancer 20 is set to about 1/2 of the reciprocating weight WREC. Therefore, for example, when the eccentricity E of the cylinder center line C0 is set to 1/2 of the distances O1 to O2, | Mzb | max0 When the amount of eccentricity E is set to 1/4 of the distances O1 to O2, | Mzb | max1 / 2 × | Mza | max becomes, and the crankshaft rotation excitation force Mzb by the balance mechanism of FIG. 2 is larger than that of the conventional example Mza. It is possible to make it extremely small.

FIG. 3 is a graph comparing the rotational excitation force Mza (broken line) of the conventional example and the rotational excitation force Mzb (solid line) by the balance mechanism of FIG. The horizontal axis represents the crank angle θ, and the vertical axis represents the crankshaft rotation Mz. The graph of Mza and Mzb in FIG. 3 is an example in which the following specific numerical values are substituted into the above-described formulas of Mza and Mzb.

WBI = 0.155 kg, R = 36 mm, RBI = 13 mm,
L = 118 mm, λ = 0.305, E = 12 mm, δ = 0.
5, g = 9.8 m / sec ² , X1 = 0 mm, Y1 = 72
mm, WREC = 0.31 kg, WROT = 0.3865 kg According to FIG. 3, the excitation moment of the rotational vibration around the crankshaft 12 is Mzb according to the present invention from 0 ° to 360 ° in the crank angle compared with the conventional example Mza. It can be seen that the absolute value is small over almost the entire region, and the absolute value at the peak has the largest reduction range, which greatly contributes to vibration reduction.

Further, when an engine was prototyped and the rotational vibration was measured, a vibration reduction of 13 to 45% was observed as compared with the conventional engine. Rotational vibration was evaluated by measuring the lateral vibration acceleration on a free vibration table as shown in FIG. 4, and dividing the value obtained by subtracting the parallel vibration component by the distance from the center of rotation. Results are shown in FIG. In FIG. 5, the solid line graph is a graph when the present invention is applied, and the dotted line graph is a conventional example. The experimental conditions are as follows.

Common conditions between new model and conventional example Connecting rod length L = 120mm Crank radius R = 34mm Primary dynamic balancer coordinate Y1 = 67.6mm Primary dynamic balancer coordinate X1 = -18mm Different conditions between new model and conventional example Eccentricity E of new model = 12 mm Conventional eccentricity E = 0 mm Next, the function of preventing the increase of the motive force in the cylinder center line direction will be described.

When the cylinder centerline C0 is eccentric from the crankshaft rotation center O1 as described above, the position of the counterbalancer center of gravity C1 is temporarily set to 1 with respect to the crankpin 15.
The position of the primary balancer center of gravity C2 is set to 18 with respect to the crankpin 15 by making it coincide with the position B1 of the phase difference of 80 °.
If the position B1 of 0 ° phase difference is made to coincide with the position B2 of symmetrical phase, the phase of the inertial force is shifted due to the deviation of the top dead center due to the eccentricity of the cylinder center line C0, and thus
The secondary inertial force remains, and a problem still remains in the cylinder center exciting force FX.

On the other hand, as in the present invention, the counter balancer center of gravity C
By displacing the center of gravity C2 of the primary and primary balancers rearward in the rotational direction, it is possible to match the phase of the primary vibration component of the reciprocating weight WREC with the counter balancer 17 and the primary balancer 20, and the vibration force in the cylinder center line direction is generated. The generation of FX can be suppressed. That is, it becomes possible to significantly reduce the rotational excitation force Mz and not increase the cylinder centerline direction excitation force FX compared to the case of FIG.

This will be described in detail. The cylinder centerline direction excitation force Fxb according to the structure of FIG. 2 and the cylinder centerline direction excitation force Fxa according to the structure of FIG. 7 are expressed as follows.

Fxb = WREC × RW ² [cos θ + {λ cos ² θ + sin θ (λ sin θ −ε) ³ + ε sin θ} / {1- (λ sin θ−ε) ² } ^3/2 ] / g + WRO
T × RW ² cos θ / g− (WROT + δWrec) × RW ² cos (θ−dθ) / g−δWREC × RW ² cos (θ−dθ1) / g FXa = WREC × RW ² cos θ / g + WROT × RW ² cos θ / g − (WROT + δWrec) × RW ² cos θ / g When the above two expressions are compared, it becomes | FXb | max | FXa | max. For the phase adjustment angles dθ, dθ1 and the overbalance rate δ in the above equation, FXa = FX is used by using an approximation equation in which the equation of FX is subjected to Fourier expansion and quadratic or higher is ignored.
It can be obtained by setting b and FYa = FYb = 0.

FIG. 6 is a comparative graph of the vibration force FX in the cylinder center line direction. FXa (broken line) is the cylinder centerline direction exciting force according to the structure of FIG. 7, FXb (solid line) is the cylinder centerline direction exciting force according to the structure of FIG. 1, FXc (phantom line) is the balancer center of gravity C1, C2. For example, position B in FIG. 1
Cylinder center line direction exciting force when they are matched with 1 and B2. That is, the balancer center of gravity C1 and C2 are respectively set to the first
When the positions B1 and B2 in the figure are matched, | FXc | m
Although the problem that ax becomes large remains, the structure of FIG. 1 can prevent the increase of | FXb | max.

FXb in the graph of FIG. 6 is an example in which the phase adjustment angle dθ of the counter balancer is set to 2 ° and the phase adjustment angle dθt of the primary balancer is set to 6.6 °.

Although both the adjustment angles dθ and dθt are different depending on the size of the engine and the like, they are set within the range of angle α1 at least as shown in FIG. The flank angle range α1 is a range α1 formed between an extension line of the connecting rod center line L1 at the top dead center and a line L2 passing through the crankshaft rotation center O1 and parallel to the cylinder center line C0. The overbalance rate δ is set to an optimum value according to the values of dθ and dθt and balanced.

(Effects of the Invention) As described above, according to the present invention, (1) The moment in the reverse direction around the crankshaft generated by the reciprocating inertia force of the eccentric piston or the like is canceled with respect to the moment around the crankshaft generated by the balancer. To work. Therefore, rotational vibration around the crankshaft is reduced. Therefore, the vibration is reduced especially when the load is low. By making the cylinder center line eccentric, it is possible to reduce the rotational vibration without providing the two primary balancers. That is, since it is not particularly necessary to provide the two primary balancers, it is possible to keep the engine compact and suppress an increase in cost.

(2) The center of gravity of the counter balancer provided on the crank web is shifted backward from the position of the phase difference of 180 ° with respect to the crank pin in the rotational direction, and the center of gravity of the primary balancer is
From the position of symmetrical phase to the above position with a phase difference of 180 ° with respect to the crank pin, since it is shifted to the rear side in the rotation direction of the primary balancer, the primary vibration of the counter balancer 17 and the primary balancer 20 and the reciprocating weight WREC. It is possible to match the phases of the components, and it is possible to prevent an increase in the cylinder centerline direction exciting force FX. That is, it becomes possible to significantly reduce the rotational excitation force Mz and not increase the cylinder centerline direction excitation force FX compared to the case of FIG.

[Brief description of drawings]

FIG. 1 is a schematic vertical sectional view of an engine to which the present invention is applied, and FIG.
Fig. 1 is a force relationship diagram of Fig. 1, Fig. 3 is a graph showing changes in rotational excitation force around a crankshaft, Fig. 4 is a schematic diagram showing a method for measuring rotational vibration, and Fig. 5 is a graph showing the result. , FIG. 6 is a graph showing a change in exciting force in the cylinder center line direction,
The figure is a vertical sectional view of a conventional example. 11 ... Cylinder, 12
...... Crank shaft, 15 ...... Crank pin, 16 ...... Crank web, 17 ...... Counter balancer, 20 ...... 1
Axis primary balancer, C0 …… Cylinder center line, O1 ……
Crankshaft rotation radius center, C1 ... Center of gravity of counter balancer, C2 ... Center of gravity of primary balancer

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Akio Amiguchi 1-1 Kawasaki-cho, Akashi-shi, Hyogo Prefecture Kawasaki Heavy Industries Ltd. Akashi Factory

Claims (1)

[Claims] 1. A 1-axis primary balancer that rotates at the same speed as the crankshaft and in the opposite direction is provided parallel to the crankshaft, and the cylinder center line is eccentric to the primary balancer side with respect to the rotation center of the crankshaft. , The center of gravity of the counter balancer provided on the crank web is shifted backward from the position of the phase difference of 180 ° with respect to the crankpin in the rotational direction, and the center of gravity of the primary balancer is described above with respect to the phase difference of 180 ° with respect to the crankpin. An engine balance mechanism characterized by being shifted rearward in the rotational direction of the primary balancer from the position of symmetrical phase to the position.