JPH0637918B2 - Balancer device for 4-cycle 2-cylinder engine - Google Patents

Description

TECHNICAL FIELD The present invention relates to a balancer device for a four-cycle two-cylinder operation engine.

(Prior Art) Among four-cycle reciprocating engines, there are only two cylinders and the operation is always performed by the two cylinders. In a cylinder control engine, for example, four cylinders are used, and two cylinders are operated under predetermined operating conditions. There is one that is designed to perform a reduced cylinder operation using only cylinders. Then, in the engine in which the two-cylinder operation is performed as described above, it is general that the combustion is performed with a 360 ° crank angle phase.

In the engine in which the two-cylinder operation is performed, the vibration of the engine, in particular, the primary vibration moment generated by combustion becomes considerably large. For this reason, in the past, Japanese Patent Laid-Open No. 60-
As disclosed in Japanese Patent No. 69344, a balancer that is rotationally driven by the crankshaft of the engine is provided, and the primary vibration moment generated by this balancer cancels (cancels) the primary vibration moment generated by combustion of the engine. Something like that is proposed.

However, while the primary vibration moment due to combustion is determined by the engine load, that is, the average effective pressure,
Since the primary vibration moment generated by the balancer utilizes the centrifugal force, it is proportional to the square of the rotation speed. Therefore, the primary vibration moment that can be canceled by this balancer is only at a specific engine speed and at a specific engine load. Therefore, conventionally, the primary vibration moment due to combustion is the most problematic operation. The balancer is set according to the range, for example, when idling.

(Problems to be solved by the invention) By the way, since the primary vibration moment generated by the balancer is based on the centrifugal force, the centrifugal force is variable, that is, at least the balancer mass or the radius of gyration is at least changed. By using one of the variable balancers as a variable balancer, it is possible to completely cancel the primary vibration moment generated by combustion over a wide operating range of the engine by the variable balancer.

However, even if such a variable balancer is used, in reality, there is a limit in completely removing the primary vibration moment generated by combustion. That is, while the normal balancer is set to have a certain specific phase with respect to the crankshaft, the phase of the primary vibration moment generated by combustion will change slightly due to changes in engine load. .

Therefore, the present invention provides a balancer device for a four-cycle two-cylinder operation engine, which is capable of appropriately responding to a phase change according to the engine load of a primary vibration moment caused by combustion in consideration of the above circumstances. To provide.

(Means and Actions for Solving Problems) In order to achieve the above-mentioned object, in the present invention, attention is paid to the point that the phase of the primary vibration moment generated by combustion is advanced as the engine load increases. Then, the phase of the balancer for canceling the primary vibration moment is advanced as the engine load increases. Specifically, two cylinders are burned with a crank angle phase of 360 ° with each other, and the primary vibration moment generated by this combustion is
In a balancer device for a four-cycle two-cylinder operation engine that is canceled by a primary vibration moment generated by a balancer rotationally driven by a crankshaft, load detecting means for detecting an engine load and a phase with respect to a crank angle of the balancer are provided. A phase adjusting means for changing the phase, and a phase controlling means for receiving the output from the load detecting means, controlling the phase adjusting means, and advancing the phase of the balancer as the engine load increases. is there.

With such a configuration, it is possible to optimally set the balancer phase by following changes in the engine load, and as a result, it is possible to more effectively reduce the primary vibration moment due to combustion. it can.

(Example) Hereinafter, an example of the present invention will be described with reference to the accompanying drawings.
In the embodiment, the phase of the balancer is adjusted in advance according to the engine load, while the turning radius of the balancer is adjusted so that the magnitude of the primary vibration moment generated by the balancer itself can be adjusted. It is shown that the optimum setting can be made according to the load and the engine speed. Further, in the embodiment, assuming that the engine is a cylinder number control engine, it is shown that two cylinders are operated during reduced cylinder operation.

Based on the above, in FIG. 1, reference numeral 1 denotes an engine, which is a reciprocating four-cycle type of four in-line four cylinders 2 (2a to 2d). For each of the cylinders 2a to 2d, air is taken in through the intake manifold 3, and from each of the cylinders 2a to 2d,
It is exhausted through the exhaust manifold 4. The second cylinder 2b of the four branch pipes 3a to 3d of the intake manifold 3
A shutter valve 5 is provided in each of the two branch pipes 3b and 3c corresponding to the third cylinder 2c, and both shutter valves 5 are opened or closed by an electromagnetic actuator 6. It is controlled to open and close. When both shutter valves 5 are open, all cylinders 2
When the operation (combustion) according to a to 2d is performed in all cylinders, and both shutter valves 5 are closed, No. 2 and No. 3 are performed.
The two cylinders numbered 2b and 2c are deactivated, and only the cylinders 2a and 2d numbered 1 and 4 are in a reduced cylinder operation, that is, a two-cylinder operation.

As shown in FIGS. 2 to 4, the engine 1 is provided with two left and right balance shafts 12 parallel to the crankshaft 11 with the crankshaft 11 interposed therebetween. A driven sprocket 13 is integrated with each one end of the balance shaft 12,
A chain 16 is provided across the both driven sprockets 13, the drive sprocket 14, and the idle sprocket 15. The drive sprocket 14 is located on the extension line of the crankshaft 11 and is connected and disconnected at the crank time 11 via an electromagnetic clutch 17, and the electromagnetic clutch 17 operates in the reduced cylinder operation described above. The balance shafts 12 are connected to each other at a predetermined timing, which will be described later.
It is driven to rotate by 1.

The balance shafts 12 are respectively provided with balancers 18 as shown in FIG. This balancer 18
In view of the fact that the radius of gyration is a variable type variable balancer, in the embodiment, the balancer pieces 18A and 18B are divided into two in the axial direction of the balance shaft 12 so as to be adjacent to each other. There is. This balancer piece 1
As shown in FIG. 2 and FIG. 6, 8A and 18B are formed in a substantially semicircular congruent shape with each other, and both balancer pieces 18A and 18B are formed on the outer peripheral surface of the balance shaft 12. It is slidably engaged with the spline portion 12a. The double spline portions 12a are formed so that their twist angles are equal and opposite to each other.

Both balancers 18A and 18B are sandwiched by a pair of first and second holding members 19 and 20 from the axial direction.
The first holding member 19 is urged toward the balancer 18 by the spring 21, and the second holding member 20 is configured as a piston rod slidably fitted in the cylinder 22, and the cylinder 22 itself. Is not relatively displaceable in the axial direction of the balance shaft 13. When the hydraulic pressure is supplied to the oil chamber 23 defined in the cylinder 22 by the second holding member 20, the second holding member 20 moves toward the balancer 18, that is, the spring 2 described above.
It is displaced in the direction of shrinking 1. The hydraulic pressure supplied to the oil chamber 23 is made constant by a constant pressure from a hydraulic pressure source by an actuator 23a composed of an electromagnetic pressure regulating valve.

As a result, both balancer pieces 1 are placed at a position balanced with the urging force of the spring 21 in accordance with the hydraulic pressure supplied to the oil chamber 23.
8, 18B are slidably changed while rotating relative to each other by the action of the double spline portion 12a. Depending on the position when this sliding change is made, both balancers 18A, 18A
The sandwiching angle θ formed by B (see FIG. 7) is changed, and the change in the sandwiching angle θ changes the radius of gyration of the combined center of gravity of the balancer pieces 18A and 18B. That is, the center of gravity position of both balancer pieces 18A and 18B is g, the composite center of gravity position is G, and the radius of gyration of the center of gravity g of one balancer piece 18A is R.
_1. Let R be the radius of gyration of the composite center of gravity G.
/ R ₁ is changed as shown in FIG. 8 by changing the included angle θ. Of course, this change of the radius gyration ratio R / R ₁ is changed according to the hydraulic pressure (oil liquid amount) supplied to the hydraulic chamber 23, as described above.

The chain angle θ of the balancers 18A and 18B is changed by rotating the balancers 18A and 18B in the same or opposite directions, but the phase of the combined center of gravity G with respect to the balance shaft 1 is unchanged.

Here, while the magnitude of the one-time excitation moment resulting from combustion is determined according to the engine load, the magnitude of the primary excitation moment due to the balancer 18 is its rotation speed (squared).
Therefore, in order to match the magnitude of the primary excitation moment due to the balancer 18 with the magnitude of the primary excitation moment due to combustion over a wide operating range of the engine, the balancer 18 has Adjusting the turning radius ratio R / R _{1 according} to the engine load and engine speed (variable)
There is a need to. FIG. 9 shows an example in which the turning radius ratio R / R ₁ of the balancer 18 is mapped so as to be optimally set according to the engine load and the engine speed, taking a certain engine as an example. As modes, seven types up to are shown. That is, embodiments are when the rotational radius ratio R / R ₁ is a 1, the rotational radius ratio aspects sequentially 1/2 increments the R / R _1,
,,,,, are shown in this order. Of course, the modes (1) to (4) are just typical examples, and the optimum conditions corresponding to the engine load and the engine speed are provided over the entire range of the region where the reduced-cylinder operation is performed (A region shown by the dashed line in FIG. 9). The radius of gyration R / R _{1 of} is finely mapped. It should be noted that it is well known how to determine the rotation radius ratio R / R _{1 according} to the engine load and the engine speed while taking the mass thereof into consideration, and the above-mentioned JP-A- Since it is also disclosed in Japanese Patent Publication No. 60-69344, further detailed description will be omitted.

Further, the balance shaft 12 has a brake mechanism 24.
Is attached. The brake mechanism 24 regulates the arbitrary rotation, that is, the reverse rotation of the balance shaft 12 during the all-cylinder operation (as described above, the electromagnetic clutch 17 is disconnected and the ballast shaft 12 is not rotationally driven during the all-cylinder operation). This is for accurately obtaining a predetermined phase relationship with the crankshaft 11 when shifting to the reduced cylinder operation in which 12 is rotationally driven. The brake mechanism 24 includes a disc 25 integrated with the balance shaft 12, and a hydraulic cylinder mechanism 23 that presses (brakes) and releases the brake (releases the brake) according to supply and discharge of hydraulic pressure (see the drawings. Is shown in simplified form), and the brakes by appropriately opening and closing electromagnetic on-off valves 29, 30 connected to the hydraulic supply / discharge pipes 27, 28 for the hydraulic cylinder mechanism 26.
The brake is released.

Now, referring to FIG. 2, the crankshaft 1 of the balancer 18
The phase variable mechanism 41 for 1, namely the portion for adjusting the advance angle will be described. The phase changing mechanism 41 includes a pair of rollers 42 and 43, a support member 44 that rotatably holds the rollers 42 and 43, and an electromagnetic actuator 45 that drives the support member 44 to move back and forth. . One of the rollers 42 and 43 is in contact with the chain 16 between the drive sprocket 14 and the one driven sprocket 13 from below. The other roller 43 is in contact with the chain 16 between the drive sprocket 14 and the other driven sprocket 13 from above. And both balance shafts 12 (both driven sprockets 13) are crankshafts 11
Therefore, when the two rollers 42, 43 are expanded by the actuator 45 from the position indicated by the solid line in FIG. 2 to the position indicated by the broken line in FIG. 2 by the actuator 45, the balance shaft is rotated. 12 (balancer 18) is advanced with respect to the crankshaft 11 (moves in the rotational direction of the balance shaft 12).

Here, the relationship between the engine load, that is, the average effective pressure and the phase angle ψ ₁ of the primary vibration moment resulting from combustion (gas pressure) is shown by taking a certain engine as an example.
It is a figure. This phase angle ψ ₁ will be described in detail later, but as is known, when the torque harmonic coefficient (only the first-order component) of the engine is a ₁ and b _{1 as} shown in FIG. 11, ψ ₁ = It becomes tan ⁻¹ (b ₁ / a ₁ ). Then, a ₁ and b ₁ are substantially unrelated to the engine speed, and become a linear function as is clear from FIG. 11, and c _{1 to} c ₄ are constants and the average effective pressure is Pe. Then, a ₁ = c ₁ Pe + c ₂ (1) b ₁ = c ₃ Pe + c ₄ (2), and thus the phase angle ψ ₁ is Becomes Then, in the above engine, since c ₁ = 0,275 c ₂ = 0,45 c ₃ = 0,75 c ₄ = 1,75, it was created based on these numerical values of c _{1 to} c ₄ . It is shown in FIG. As is clear from FIG. 12, it is understood that the phase angle ψ ₁ of the primary vibration moment due to combustion becomes more orthogonal as the engine load (average effective pressure Pe) increases.

In FIG. 1, FIG. 2 and FIG. 4, 51 is a control unit composed of a microcomputer.
Signals from the sensors 52 to 56 are input to the control unit 51, while the control unit 51 outputs actuators 6 (for changing the number of cylinders), 23a (for adjusting the radius of rotation of the balancer 18), and 45 (for adjusting the balancer phase). ), The clutch 17 (for driving the balancer 18), and the opening / closing valves 29, 30 (for braking). Of the sensors 52 to 56, the sensor 52 (see FIG. 1) is a water temperature sensor that detects the cooling water temperature of the engine. The sensor 53 (see FIG. 1) is
It is a load sensor that detects the intake negative pressure immediately downstream of the throttle valve 57, that is, the engine load. The sensor 54 (see FIG. 1) is a rotation speed sensor that detects the engine rotation speed (the rotation speed of the crankshaft 11). The sensor 55 (see FIG. 4) is a phase sensor for detecting the phase of the balancer 18, and in the embodiment, is constituted by, for example, a pickup, and detects the phase of the driven sprocket 13. The sensor 56 (see FIG. 4) is the crankshaft 11
A phase sensor for detecting the phase of the cam shaft. In the embodiment, for example, a sprocket 58 (see FIG. 4) for driving the cam shaft, which is constituted by a pickup and is integrated with the crank shaft 11,
The phase of is detected.

The control unit 51 described above controls the number of cylinders for switching between the all-cylinder operation and the reduced-cylinder operation based on the flowchart shown in FIG. 10, and also (the combined center of gravity G of the balancer 18 during the reduced-cylinder operation. ) The radius of gyration and its phase are controlled. In FIG. 10, first, at step S1, the current cooling water temperature T of the engine 1
W, intake negative pressure (engine load) Pb, engine speed N
a is read. After that, in step S2, S3 or S4, each of the values TW, Pb, Na is compared with a predetermined set value To, Po or No, and if TW ≥To is not satisfied, the process directly returns to step S1 or Pb ≥P.
Either when not o or when Na ≤ No, 1
If it falls under one of the two, step S
Move to 9. In step S9, the actuator 6 for switching the reduced cylinder operation is turned off, that is, the shutter valve 5 is opened. Next, in step S8, the variable balancer 18
Adjustment actuators are turned off, that is, the balance shaft 12 is braked at least by the brake mechanism 24, the adjustment of the phase variable mechanism 41 is stopped, and the radius of gyration of the balancer 18 (the composite gravity center G thereof) is adjusted. To stop.

On the other hand, in the determination of steps S2, S3, and S4, TW
When all of the conditions of ≧ To, Pb ≧ Po, and Na ≦ No are satisfied, it is the condition of the reduced cylinder operation.
In step S5, the actuator 6 is turned on to close the shutter valve 5 (start of reduced cylinder operation). Next, in step S6, the optimum turning radius (turning radius ratio R / R ₁ ) of (the combined center of gravity G of) the balancer 18 is calculated from the map as shown in FIG. 7 according to the engine load and the engine speed, and The optimum phase of the balancer 18 is calculated according to the map shown in FIG. 12 according to the engine load. Then, in step S7, the oil chamber 23 is adjusted to the set value calculated in step S6.
The hydraulic pressure supplied to the balancer is adjusted (the radius of rotation of the balancer is adjusted), and the actuator 45 is adjusted (the phase of the balancer 18 is adjusted).

The cancellation of the primary vibration moment by adjusting the radius of gyration and the phase of the balancer 18 as described above is schematically shown in FIGS. 13 and 14. That is, in FIG. 13, the α line indicated by the solid line is the primary vibration moment resulting from combustion, and the β line indicated by the broken line is the primary vibration moment generated by the balancer 18. And, the solid line 1 in FIG. 14 shows the γ-rays as the exciting moment before the primary exciting moment is removed, and the δ-line as shown by the broken line is the balancer 1
8 is the excitation moment after canceling the primary excitation moment. From the comparison between the γ line and the δ line in FIG. 14, the effect of canceling the primary excitation moment by the balancer 18 (reduction of excitation moment) can be easily understood.

Next, the reason why the balancer 18 must be advanced in accordance with the engine load will be described in detail. First, in FIG. 2, the Y axis is an axis that passes through the axis of the crankshaft 11 and extends in the cylinder axis direction, and the axis orthogonal to the Y axis is the X axis. To (x
₁ , y ₁ ) and (x ₂ , y ₂ ). Further, both balancers 18 are assumed to be rotated in the reverse direction with respect to the crankshaft 11, and are at the same position in the axial middle portion of the crankshaft 11 in the Z-axis direction (the direction perpendicular to the paper surface). The crank angle of the crankshaft 11 is θ (angle of rotation from the top dead center position), the mass of the balancer 18 is m, and the radius of rotation is r.
(Corresponding to R in FIG. 7), the angular velocity is ω, the deviation angle (phase) of the one balancer 18 with respect to the crank angle θ is α ₁ , the deviation angle of the other balancer 18 with respect to the crank angle θ is α ₂ , the engine Let V be the total displacement of Also,
As described above, the torque harmonic coefficients are set to a ₁ and b ₁ (see FIG. 11).

Based on the above, the primary excitation moment MGAS resulting from combustion is MGAS = V (a ₁ cos θ + b ₁ sin θ) / 2 (2). In addition, the primary vibration moment MBAL by both balancers 18 is Furthermore, in order to statically balance both balancers 18 themselves, m ₁ r ₁ cos α ₁ + m ₂ r ₂ cos α ₂ = 0 (6) m ₁ r ₁ cos α ₁ + m ₂ r ₂ sin α ₂ = It is necessary to satisfy 0 (7).

Then, in order to cancel the primary excitation moment MGAS due to combustion with the primary excitation moment MBAL generated by the balancer 18, it is sufficient to make both MGAS and MBAL equal, and since this MGAS = MBAL, (4 ), From equation (5), −x ₁ m ₁ r ₁ ω ² cos α ₁ −x ₂ m ₂ r ₂ ω ² cos α ₂ + y ₁ m ₁ r ₁ ω ² sin α ₁ + y ₂ m ₂ r ₂ ω ^{_{2 sin α 2 = V · a}} 1/2 (8) -x 1 m 1 r 1 ω 2 sin α 1 -x 2 m 2 r 2 ω 2 sin α 2 -y 1 m 1 r 1 ω 2 cos α 1 ^{_{-y 2 m 2 r 2 ω 2}} cos α 2 = V · b 1/2 (9) (6), (7) equation (8), are substituted into equation (9), cos alpha ₂₄ section And solving for sin α ₂ term, From equations (10) and (11) Here, assuming that the coordinate system is y ₁ = y ₂ and x ₁ = x ₂ for simplification, Equation (12) is expressed as α ₂ = tan ⁻¹ (b ₁ / a ₁ ) (13) and (6 ), From equation (7), tan α ₁ = tan α ₂ (14) Therefore, from this equation (14), α ₁ = α ₂ ± nπ (n = 1, 2, ...) , (9), one α ₂ is determined for α ₁ . Therefore, tan α ₁ = (tan α ₂ ) tan ⁻¹ (b ₁ / a ₁ ) (15) Becomes This equation (17) clearly shows that it increases as the engine load (average effective pressure Pe) increases, as shown in FIG.

As described above, by advancing the phases of the balancer 18 by α ₁ and α ₂ according to the engine load, the phase change according to the engine load of the primary vibration moment caused by combustion is tracked, and The secondary vibration moment can be removed by the balancer 18. By the way, when the average effective pressure Pe is in the range of −5 kg / cm ^{2 in} a reduced cylinder operation, α ₁ (same for α ₂ ) is −99, 93 ° to 108,
It may be changed within a range of 8,43 °, such as 36 °. It should be noted that normally, a reference phase (α ₁ , α ₂ ) for a certain specific engine load may be set, and the reference phase may be advanced or retarded in accordance with the change in the engine load. This is the same for the adjustment of the radius of gyration of the balancer 18 (however, with respect to the radius of gyration, the reference engine load and engine speed are specified).

(Effects of the Invention) As is apparent from the above description, even if the phase of the primary vibration moment resulting from combustion fluctuates according to the fluctuation of the engine load, the phase of the balancer also fluctuates accordingly. Since it is changed, the primary vibration moment resulting from this combustion can be removed with higher accuracy, and a preferable one can be obtained in terms of reducing engine vibration.

[Brief description of drawings]

FIG. 1 shows an embodiment of the present invention, and is a simplified partial cross-sectional plan view of a variable cylinder number engine. FIG. 2 is a simplified diagram showing the relationship between the crankshaft, the balancer, and the phase adjusting means of the balancer. FIG. 3 is a partial side sectional view of a balancer driving portion of the engine shown in FIG. FIG. 4 is an enlarged cross-sectional view of the main part of FIG. FIG. 5 is a side sectional view showing a portion of the balance shaft and the balancer. FIG. 6 is a view of one of the balancer pieces of the split type balancer as seen from the axial direction of the balance shaft. FIG. 7 is an explanatory view of a balancer composed of two balancer pieces as viewed in the axial direction. FIG. 8 is a graph showing the relationship between the gripping angle of the balancer shown in FIG. 7 and the turning radius ratio. FIG. 9 is a graph (map) showing an example of the optimum balance radius ratio (angle of inclusion) of the balancer corresponding to the engine speed and the engine load. FIG. 10 is a flowchart showing a control example of the present invention. FIG. 11 is a graph showing an example of the torque harmonic coefficient of the first-order component. FIG. 12 is a graph showing an example of the relationship between the average effective pressure and the phase of the primary vibration moment resulting from combustion. FIG. 13 is a graph showing the relationship between the primary vibration moment due to combustion and the primary vibration moment due to the balancer. FIG. 14 is a graph showing the states of the exciting moment before and after the removal of the primary exciting moment due to combustion. 1: Engine 2: Cylinder 2b, 2c: Inactive cylinder (during reduced cylinder operation) 5: Shutter valve (for switching the number of cylinders) 6: Actuator (for switching the number of cylinders) 11: Crankshaft 12: Balance shaft 12a: Double spline Part 13: Driven sprocket 14: Drive sprocket 16: Chain 17: Electromagnetic clutch (for balancer drive) 18: Balancer 18A, 18B: Balancer piece 22: Cylinder (for changing balancer rotation radius ratio) 23: Oil chamber 23a: Actuator ( Hydraulic pressure adjustment) 41: Phase changing mechanism 42, 43: Roller 44: Support member 45: Actuator (for phase change) 51: Control unit 53: Sensor (for engine load detection)

Claims (1)

[Claims] 1. Two cylinders are burned with a crank angle phase of 360 ° with each other, and a primary vibration moment generated by this combustion is canceled by a primary vibration moment generated by a balancer rotationally driven by a crankshaft. In the balancer device for the four-cycle two-cylinder operation engine, the load detecting means for detecting the engine load, the phase adjusting means for changing the phase with respect to the crank angle of the balancer, and the phase adjusting means for receiving the output from the load detecting means. And a phase control means for advancing the phase of the balancer as the engine load increases, and a balancer device for a four-cycle two-cylinder operation engine.