RU2671666C1 - Balancing device for internal combustion engine - Google Patents

Abstract

FIELD: internal combustion engines.

SUBSTANCE: balancing device for an internal combustion engine includes a crankshaft and a balancing shaft. Crankshaft includes an eccentric crankshaft load. Balancing shaft includes an eccentric weight of the balancing shaft. Point of connection with the crankshaft, deflected from the main shaft of the crankshaft, and the point of connection with the balancing shaft, deflected from the axial shaft of the balancing shaft, is connected to the connecting rod. Crankshaft connection mechanism, which ensures the relative rotation of the crankshaft and the connecting rod, is provided at the point of connection to the crankshaft. Coupling mechanism with a balancing shaft, which ensures the relative rotation of the balancing shaft and the connecting rod, is provided at the point of coupling with the balancing shaft. Guide section guides the movement of the connecting rod the way the balancing shaft rotates in the opposite direction relative to the crankshaft.

EFFECT: technical result is to reduce the size and weight of the internal combustion engine.

20 cl, 69 dwg

Description

State of the art

Embodiments of the present invention relate to a balancing device for an internal combustion engine, and in particular, relate to a balancing device suitable for mounting on a single cylinder or two cylinder four stroke internal combustion engine.

State of the art

The balancing device is generally mounted on a reciprocating internal combustion engine. During operation of the internal combustion engine, an inertia force arises due to the movement of the piston. The balancing device is configured to generate a vibration-causing force to balance the vibration caused by the inertia force. If the balancing device appropriately balances the vibration, an internal combustion engine with excellent smoothness can be implemented.

JP 2010-169045 A discloses a balancing device for mounting on a four cylinder internal combustion engine. The balancing device has a balancing shaft to which an eccentric load is connected. The balancing shaft is connected to the crankshaft through gears with an unequal speed of rotation. When the crankshaft rotates during operation of the internal combustion engine, the balancing shaft rotates through gears with an unequal speed of rotation.

At this time, the eccentric load, which is attached to the balancing shaft, periodically generates a vibration-causing force in accordance with the angular velocity and angular acceleration of the balancing shaft. The angular velocity and angular acceleration of the balancing shaft vary with profiles corresponding to the characteristics of gears with an unequal rotation speed. According to JP 2010-169045 A, gears with an unequal speed of rotation are formed in such a way that a large vibration-causing force is generated at the angle of rotation of the crankshaft, at which the vibration to be balanced is large. Therefore, according to the above-described conventional balancing device, vibration of the internal combustion engine can be effectively suppressed, and excellent smooth operation can be realized.

The problem to which the invention is directed

However, in the balancing device described in JP 2010-169045 A, it is necessary to transmit the rotation of the crankshaft to the balancing shaft by means of gears. Therefore, in an internal combustion engine in which the distance between the crankshaft and the balancing shaft is large, the gears do not allow to contain the increase in size. As a result, the size of the balancing device increases, and a situation arises when it is difficult to reduce the size and weight of the internal combustion engine.

Embodiments of the present invention are implemented in order to solve the problem as described above, and have the goal of arranging a balancing device that can effectively balance the vibration of the internal combustion engine without the difficulty of reducing the size and weight of the internal combustion engine.

SUMMARY OF THE INVENTION

To achieve the above objective, a first aspect of an embodiment of the present invention is a balancing device for an internal combustion engine, comprising:

a crankshaft that rotates with the main shaft of the crankshaft as a rotational shaft; and

a balancing shaft that rotates with the axial shaft of the balancing shaft parallel to the main shaft of the crankshaft as a rotational shaft,

wherein the crankshaft includes an eccentric load of the crankshaft, which sets the center of gravity of the crankshaft to be eccentric relative to the center of the main shaft of the crankshaft, and

the balancing shaft includes an eccentric load of the balancing shaft, which sets the center of gravity of the balancing shaft to be eccentric relative to the center of the axial shaft of the balancing shaft,

wherein the balancing device further comprises:

a connecting rod, which connects the connection point with the crankshaft located on the crankshaft in a position deviated from the center of the main shaft of the crankshaft, and the connection point with the balancing shaft located on the balancing shaft in a position deviated from the center of the axial shaft of the balancing shaft;

a mechanism for connecting to the crankshaft, which provides relative rotation of the crankshaft and connecting rod with a connection point with the crankshaft as the center of rotation;

a mechanism for connecting to the balancing shaft, which provides relative rotation of the balancing shaft and connecting rod with a connection point with the balancing shaft as the center of rotation; and

a guide section that guides the movement of the connecting rod in such a way that the balancing shaft rotates in the opposite direction relative to the direction of rotation of the crankshaft.

A second aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a first aspect explained above, further comprising:

a mechanism for adjusting the connection points, which allows at least one of the connection point with the crankshaft and the connection point with the balancing shaft to move in the direction of the radius of rotation of at least one of the connection point with the crankshaft and the connection point with the balancing shaft; and

the sliding part located at one point of the connecting rod,

wherein the guide section controls the movement of the sliding part to a rectilinear movement in the direction from the side of the main shaft of the crankshaft to the side of the axial shaft of the balancing shaft and the rectilinear movement in the opposite direction with respect to the direction.

A third aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a first aspect explained above, further comprising:

a mechanism for adjusting the connection points, which allows at least one of the connection point with the crankshaft and the connection point with the balancing shaft to move in the direction of the radius of rotation of at least one of the connection point with the crankshaft and the connection point with the balancing shaft;

however, the guide section can rotate in a plane identical to the connecting plane of the connecting rod with a position superimposed on the connecting rod as the center, and holds the connecting rod with the possibility of sliding in the direction of the center line of the connecting rod.

A fourth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a first aspect explained above, further comprising:

a restrictive part located at a midpoint between the connection point with the crankshaft and the connection point with the connecting rod balancing shaft,

the distance between the connection point with the crankshaft and the connection point with the balancing shaft is equal to the distance between the main shaft of the crankshaft and the axial shaft of the balancing shaft,

the distance between the center of the main shaft of the crankshaft and the connection point with the crankshaft is equal to the distance between the center of the axial shaft of the balancing shaft and the connection point with the balancing shaft, and

the guide section includes a guide on the side of the balancing shaft, which prevents the restriction part from shifting in the direction of rotation identical to the direction of rotation of the connection point with the crankshaft in a position in which the restriction part is closest to the axial shaft of the balancing shaft, and the guide on the crankside side shaft, which prevents the restrictive part from shifting in the direction of rotation identical to the direction of rotation of the connection point with the crankshaft in The position in which the restrictive part is closest to the main shaft of the crankshaft.

A fifth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a second aspect explained above, in which:

the crankshaft is used by the offset crank method, in which the center of the main shaft of the crankshaft is set in a position that is offset by a fixed value from the center line of the reciprocating motion of the piston, and

the balancing shaft and the guide section are arranged so that at least one of the center of the main shaft of the crankshaft and the center of the axial shaft of the balancing shaft is placed in a position that is shifted by a fixed value from the axial line of rectilinear motion.

A sixth aspect and an embodiment of the present invention is a balancing device for an internal combustion engine according to a fifth aspect explained above, in which the center line of the crankshaft to the balancing shaft connecting the center of the main shaft of the crankshaft and the center of the axial shaft of the balancing shaft is shifted by a fixed value from the centerline of rectilinear motion.

A seventh aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a fifth aspect explained above, in which:

the center of the main shaft of the crankshaft is located on the axial line of rectilinear motion, and

the center axis of the balancing shaft is shifted by a fixed value from the center line of the rectilinear motion.

An eighth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a fifth aspect explained above, wherein:

the center of the axial shaft of the balancing shaft is located on the axial line of rectilinear motion, and

the center of the main shaft of the crankshaft is shifted by a fixed value from the center line of the rectilinear motion.

A ninth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a fifth aspect explained above, in which:

the center axis of the balancing shaft is shifted by a fixed value to one side of the axial line of rectilinear movement, and

the center of the main shaft of the crankshaft is shifted by a fixed value to the other side of the axial line of rectilinear motion.

A tenth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a third aspect explained above, in which:

the crankshaft is used by the offset crank method, in which the center of the main shaft of the crankshaft is set in a position that is shifted by a fixed value from the reciprocating motion of the piston, and

the center of rotation of the guide section is shifted by a fixed value from the center line "crankshaft - balancing shaft" connecting the center of the main shaft of the crankshaft and the center of the axial shaft of the balancing shaft.

An eleventh aspect and an embodiment of the present invention are a balancing device for an internal combustion engine according to any one of the first to tenth aspects explained above, wherein:

the connection point with the crankshaft is provided on the side identical to the center of gravity of the eccentric load of the crankshaft relative to the center of the main shaft of the crankshaft, and

the connection point with the balancing shaft is provided on the side identical to the center of gravity of the eccentric load of the balancing shaft relative to the center of the axial shaft of the balancing shaft.

A twelfth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to any one of the first to tenth aspects explained above, wherein:

the connection point with the crankshaft is provided on the opposite side with respect to the center of gravity of the eccentric load of the crankshaft relative to the center of the main shaft of the crankshaft, and

the connection point with the balancing shaft is provided on the opposite side with respect to the center of gravity of the eccentric load of the balancing shaft relative to the center of the axial shaft of the balancing shaft.

A thirteenth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to any one of the first to twelfth aspects explained above, further comprising:

a torque application mechanism that applies torque to the balancing shaft in the opposite direction with respect to the direction of rotation of the crankshaft.

A fourteenth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to a thirteenth aspect explained above, wherein:

the torque application mechanism includes a cam, which is provided on the balancing shaft, and a spring element, which is compressed by pressing by the cam, and

the cam is formed in such a way that it presses the spring element during the movement of the connecting rod to the side of the axial shaft of the balancing shaft with the rotation of the balancing shaft and receives a torque in the opposite direction from the spring element in the position in which the axial line of the connecting rod is superimposed on the axial shaft of the balancing shaft.

A fifteenth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to any one of the first to fourteenth aspects explained above, wherein the balancing device is mounted on a single cylinder or four stroke two cylinder internal combustion engine.

A sixteenth aspect and an embodiment of the present invention are a balancing device for an internal combustion engine according to any one of the first to fifteenth aspects explained above, wherein:

the connecting rod is positioned so that it is inclined relative to the center line of the reciprocating motion of the piston at the top dead center and bottom dead center of the internal combustion engine,

the eccentric load of the crankshaft has a center of gravity in the region that is on the opposite side relative to the connection point with the crankshaft, with the axial line of the crankshaft, which passes through the center of the main shaft of the crankshaft and is parallel with the axial line of the piston between them, in the situation in the dead point and

the eccentric load of the balancing shaft has a center of gravity in the region that is on the opposite side with respect to the connection point with the balancing shaft, with the axial line of the balancing shaft, which passes through the center of the axial shaft of the balancing shaft and is parallel with the axial line of the piston between them, in the situation in the upper dead point.

A seventeenth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to the sixteenth aspect explained above, in which:

the eccentric load of the crankshaft has a center of gravity and weight with an absolute value that balances the resulting force from the vibration-induced force caused by the connecting rod of the internal combustion engine, the part that causes the vibration caused by the piston of the internal combustion engine, and the part that causes the vibration caused by the connecting rod ,

the eccentric load of the balancing shaft has a center of gravity and weight with an absolute value that balances the remainder of the vibration causing force caused by the piston of the internal combustion engine and the rest of the vibration causing force caused by the connecting rod, and

said parts of the vibrational force and said remaining parts of the vibrational force are equal.

An eighteenth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to the seventeenth aspect explained above, in which:

the balancing shaft is connected to the connecting rod at one end of the balancing shaft, and

From the weight of the eccentric load of the balancing shaft, the weight to balance the remaining part of the vibration-induced force caused by the connecting rod is reflected more significantly in the vicinity of one end compared to the vicinity of the other end of the balancing shaft.

A nineteenth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to any one of the first to eighteenth aspects explained above, wherein:

the connecting rod has a bearing on the crankshaft side on the crankshaft side,

the crankshaft coupling mechanism has an eccentric shaft on the crankshaft side, which is rotatably held by a bearing on the crankshaft side,

an eccentric shaft on the side of the crankshaft is attached to the main shaft of the crankshaft so that the eccentric point of the crankshaft, which deviates by a fixed value from its center, coincides with the center of the main shaft of the crankshaft, and

the center of the eccentric shaft on the crankshaft side configures the connection point with the crankshaft.

A twentieth aspect of an embodiment of the present invention is a balancing device for an internal combustion engine according to any one of the first to nineteenth aspects explained above, wherein:

the connecting rod has a bearing on the side of the balancing shaft on the side of the balancing shaft,

the balancing shaft connection mechanism has an eccentric shaft on the balancing shaft side, which is rotatably held by a bearing on the balancing shaft side,

an eccentric shaft on the side of the balancing shaft is attached to the axial shaft of the balancing shaft so that the eccentric point of the balancing shaft, which deviates by a fixed value from its center, coincides with the center of the axial shaft of the balancing shaft, and

the center of the eccentric shaft on the balancing shaft side configures the connection point with the balancing shaft.

Advantages of Embodiments

According to a first aspect of an embodiment of the present invention, the crankshaft and the balancing shaft rotate in opposite directions relative to each other. At a time when the crankshaft and balancing shaft both rotate one revolution, the phases of the center of gravity of the crankshaft and the center of gravity of the balancing shaft coincide two times. Later in this document, the direction connecting two points that coincide with each other is called the "Y axis direction", and the direction perpendicular to the Y axis is called the "X axis direction." Further, a vibration causing force caused by the eccentric load of the crankshaft is referred to as a “vibration causing crankshaft”, and a vibration causing force caused by the eccentric load of the balancing shaft is referred to as “vibration causing the balancing shaft”. During the rotation of the crankshaft and the balancing shaft in opposite directions, the component along the X axis causing the vibration of the crankshaft and the component along the X axis causing the vibration of the balancing shaft act to balance each other. Meanwhile, the components along the Y axis of these vibration-causing forces are combined and amplified. Therefore, according to the first aspect explained above, the vibration causing force can be generated mainly in the Y axis direction. In the internal combustion engine, with the reciprocating motion of the piston, the inertia force that causes the vibration arises in the reciprocating direction. According to the first aspect explained above, by matching the direction of the Y axis with the direction of reciprocating motion, the inertia of the piston can be balanced by the resultant force from the vibration causing the crankshaft and the vibration causing the balancing shaft.

In the first aspect explained above, the eccentric load of the crankshaft rotates with the crankshaft. Consequently, the component along the Y axis of the vibration causing the crankshaft changes in the form of a sine wave with a change in the angle of rotation of the crankshaft. Meanwhile, the eccentric load of the balancing shaft rotates in the opposite direction relative to the crankshaft through the connecting rod. In this case, the rotation of the eccentric load of the balancing shaft necessarily becomes rotation at an unequal rotation speed, when the rotation of the crankshaft is rotation at an equal rotation speed. The y-axis component of the vibration causing vibration of the balancing shaft shows a change in the shape of the distorted sine wave relative to a change in the angle of rotation of the crankshaft.

The inertia caused by the reciprocating motion of the piston shows a change in the shape of the sine wave relative to the rotation of the crankshaft when the ratio of the length of the connecting rod 1c to the crank radius rc, i.e. the connecting rod ratio 1c / rc is infinity. With a practical connecting rod coefficient of 1c / rc, the inertia shows a change in the shape of the distorted sine wave relative to the change in the angle of rotation of the crankshaft. According to the first aspect explained above, the y-axis component of the vibration causing the balancing shaft changes to a distorted sine wave shape, the resulting force from the vibration causing the crankshaft and the vibration of the balancing shaft can be matched with the inertia caused by the reciprocating piston , with high accuracy. Therefore, according to the first aspect explained above, vibration of the internal combustion engine can be effectively suppressed.

In addition, the first aspect of an embodiment of the present invention can realize the above effect by means of a connecting rod and a guide section without using gears. The connecting rod and guide section can be formed in such a way that they are lighter and fit in a small space compared to gears. Therefore, according to the first aspect explained above, the vibration of the internal combustion engine can be effectively balanced without the difficulty of reducing the size and weight of the internal combustion engine.

According to a second aspect explained above, the position of the sliding part on the connecting rod is limited to any one of the points in a rectilinear movement permitted by the guide section. Hereinafter, the direction of the straight line is called the "y-axis", and the direction orthogonal to the y-axis is called the "x-axis". When the crankshaft rotates, the connection point with the crankshaft changes the position in the x-axis direction, as well as the position in the y-axis direction. The x coordinate of the sliding part is limited, and therefore, when the connection point with the crankshaft moves in the positive direction along the x axis, the connection point with the balancing shaft necessarily moves in the negative direction along the x axis. Additionally, when the direction of displacement of the connection point with the crankshaft changes to the negative direction along the x axis from the positive direction along the x axis, the direction of the offset of the points of connection to the crankshaft changes to the positive direction along the x axis from the negative direction along the x axis. In this case, the connection point with the balancing shaft always moves in a direction identical to the direction of the connection point with the crankshaft with respect to the direction along the y axis. As a result, the balancing shaft rotates in the opposite direction relative to the crankshaft.

In the third aspect explained above, the connection point with the crankshaft and the connection point with the balancing shaft are displaced in the same direction in the y direction, while in the direction along the x axis, the connection point with the crankshaft and the connection point with the balancing shaft are opposed directions, as in the case of the second aspect. Therefore, according to the third aspect explained above, the balancing shaft after rotation of the crankshaft can rotate in the opposite direction relative to the crankshaft. Additionally, in the third aspect explained above, the BS / CS ratio between the distance from the connection point with the balancing shaft to the guide section (hereinafter referred to as the "BS distance") and the distance from the connection point with the crankshaft to the guide section ( hereinafter referred to as “CS distance”) changes with the rotation of the crankshaft. In the second aspect explained above, the ratio is always constant. Based on the principle of the lever, as the above ratio increases more, the change in the angle of rotation of the balancing shaft, accompanying the change in the angle of rotation of the crankshaft, becomes large. Therefore, according to the third aspect explained above, a vibration-inducing force for balancing the piston inertia force can be provided with a profile different from the case of the second aspect explained above.

According to the fourth aspect explained above, the restrictive part of the connecting rod approaches the balancing shaft to the greatest extent in a situation in which the center of the main shaft of the crankshaft, the connection point with the crankshaft, the restriction part, the center of the axial shaft of the balancing shaft and the connection point with the balancing shaft are combined in one straight line. Later in this document, this position is called the "first point of change." At the first point of change, the axial force of the connecting rod, which acts on the connection point with the balancing shaft, does not generate torque. Therefore, if there is no restriction in the direction of movement, the connection point with the balancing shaft can rotate in any direction from the first point of change with a change in the angle of rotation of the crankshaft. When the connection point with the balancing shaft is shifted in a direction identical with the direction of rotation of the crankshaft, the balancing shaft rotates in a direction identical to the direction of the crankshaft. In the fourth aspect explained above, displacement in the above direction is prevented by the restrictive portion of the connecting rod and the guide on the side of the balancing shaft. Therefore, when the angle of rotation of the crankshaft changes from the above situation, the connection point with the balancing shaft is shifted in the opposite direction relative to the direction of rotation of the crankshaft. When the connection point is outside the first change point, the axial force of the connecting rod, which acts on the connection point with the balancing shaft, generates a torque. Consequently, the balancing shaft continues reverse rotation with the rotation of the crankshaft. When the crankshaft rotates 180 degrees from the state of the first change point, a situation is formed in which the center of the main shaft of the crankshaft, the connection point with the crankshaft, the restrictive part, the center of the axial shaft of the balancing shaft and the connection point with the balancing shaft are aligned in a straight line, in a state in which the restrictive part of the connecting rod is closest to the main shaft of the crankshaft. Later in this document, the position is called the "second point of change." At the second point of change, the displacement of the restrictive part is controlled by a guide on the side of the crankshaft. As a result, at the second change point, the connection point with the balancing shaft is guided in the opposite direction relative to the direction of rotation of the crankshaft. The above operation is repeated, whereby the balancing shaft can also continue to rotate in the opposite direction relative to the crankshaft through the connecting rod by the fourth aspect explained above.

According to any one of the fifth to ninth aspects explained above, the crankshaft is used in an offset crank manner such that the inertia caused by the movement of the piston from the top dead center to the bottom dead center and the inertia caused by the movement of the piston from the bottom dead center to top dead center, become asymmetric. If the sliding part of the connecting part must reciprocate on the center line “crankshaft - balancing shaft”, which connects the center of the main shaft of the crankshaft and the center of the axial shaft of the balancing shaft, the balancing shaft must exhibit symmetrical angular velocity profiles during passage from the upper side dead center to the side of bottom dead center and in a process that is the opposite of it. Thus, the vibrational forces that must be generated by the eccentric load of the balancing shaft in an external manner and in the opposite way become symmetrical. In contrast, in any of the fifth to ninth aspects explained above, the rectilinear movement of the sliding part is directed to a straight line that does not coincide with the center line "crankshaft - balancing shaft". In this case, a distortion occurs in the profile of the angular velocity of the balancing shaft, and the forces causing vibration, which are generated externally and in the opposite way by the eccentric load of the balancing shaft, become asymmetric. Therefore, according to any one of the fifth to ninth aspects explained above, vibrational forces can be generated that become asymmetric in the external and reverse ways, and the inertia generated by the piston under the condition of the shifted crank can be properly balanced.

According to the tenth aspect explained above, the inertia forces that are generated by the piston in an external and reverse manner are asymmetric, as in the case of the ninth aspect explained above. In the configuration in which the rotating guide section holds the connecting rod, if the center of rotation is set on the center line “crankshaft - balancing shaft”, the angular velocity profiles of the balancing shaft in an external and reverse way become symmetrical, and as a result, vibration forces generated by the eccentric load balancing shaft in an external way and the reverse way, also become symmetrical. In contrast, if the center of rotation of the guide section is outside the center line of the “crankshaft - balancing shaft", vibrational forces that are generated by the eccentric load of the balancing shaft in an external manner and in the opposite way become asymmetric. Therefore, according to the tenth aspect explained above, the inertia generated by the piston can be properly balanced under the condition of a shifted crank.

According to the eleventh aspect explained above, by synchronizing the rotation phase of the eccentric load of the crankshaft and the rotation phase of the eccentric load of the balancing shaft with each other, a vibration-causing force that changes along the desired profile can be generated in the direction along the Y axis.

According to the twelfth aspect explained above, by synchronizing the rotation phase of the eccentric load of the crankshaft and the rotation phase of the eccentric load of the balancing shaft with each other, a vibrational force that varies along a profile different from the profile realized in the eleventh aspect explained above can be formed in Y direction

In the thirteenth aspect explained above, torque is provided to the balancing shaft through a connecting rod. In this configuration, at the point of change at which the connecting rod center line is superimposed on the center of rotation of the balancing shaft, the connecting rod axial force does not apply torque to the balancing shaft. Therefore, if the external force that is applied to the balancing shaft is only the axial force of the connecting rod, the balancing shaft must be in a state in which the balancing shaft can rotate in normal and reverse directions at the point of change. In the thirteenth aspect explained above, a reverse torque is applied to the balancing shaft by a torque application mechanism. Therefore, according to the thirteenth aspect explained above, the balancing shaft can continue to stably rotate in the opposite direction with respect to the direction of rotation of the crankshaft.

According to the fourteenth aspect explained above, a suitable torque can be applied to the balancing shaft at the point of change by means of a cam and a spring element.

According to the fifteenth aspect explained above, an internal combustion engine includes one piston that works separately, or two pistons that work in the same phase. In an internal combustion engine, pistons do not mutually balance the inertia forces of the reciprocating movements. According to the fifteenth aspect explained above, the vibration of the internal combustion engine can be appropriately suppressed by the vibration causing force generated by the balancing device.

According to the sixteenth aspect explained above, in the situation of top dead center, the piston and the connecting rod generate vibrational forces in the support direction, which runs along the center line of the reciprocating motion of the piston. In this case, the connecting rod applies a force causing vibration to the eccentric load of the crankshaft in the first inclination direction, which extends almost from the center of the main shaft of the crankshaft to the connection point with the crankshaft, while also applying a vibration force to the eccentric load of the balancing shaft in the second direction inclination, which runs almost from the center of the axial shaft of the balancing shaft to the point of connection with the balancing shaft. The resulting force from the above vibrational forces has components to the first tilt direction and the second tilt direction, in addition to the component to the above reference direction. In the sixteenth aspect explained above, the center of gravity of the eccentric load of the crankshaft is provided on the opposite side of the connection point with the crankshaft, with the center line of the crankshaft between them. According to the center of gravity, the vibration causing force component in the first inclination direction can be balanced in addition to the vibration causing force component in the reference direction. Additionally, in the sixteenth aspect explained above, the center of gravity of the eccentric load of the balancing shaft is provided on the opposite side of the connection point with the balancing shaft, with the axial line of the balancing shaft between them. According to the center of gravity, the vibration causing force component in the second inclination direction can be balanced in addition to the vibration causing force component in the reference direction. At the bottom dead center of an internal combustion engine, balancing the vibration-causing forces occurs on the basis of a similar principle. Therefore, according to the sixteenth aspect explained above, vibration-causing forces that are respectively generated by the piston, connecting rod and connecting rod can be properly balanced.

According to the seventeenth aspect explained above, at the top dead center and bottom dead center of the internal combustion engine, the vibration causing forces caused by the eccentric load of the crankshaft and the eccentric load of the balancing shaft can be balanced with the vibration causing forces caused by the connecting rod, piston and connecting rod. Additionally, in a situation with the exception of top dead center and bottom dead center, the vibrational forces produced by the connecting rod and the eccentric load of the balancing shaft can be balanced with the vibrational force caused by the eccentric load of the crankshaft. Therefore, according to the seventeenth aspect explained above, vibration-causing forces caused by the individual elements can always be balanced preferably.

According to the eighteenth aspect explained above, the vibration causing force caused by the connecting rod is introduced at one end of the balancing shaft. The eccentric load of the balancing shaft, located for the balancing shaft, can balance the above vibration-causing force by means of a weight significantly reflected in the vicinity of one end. As the input point of the vibration causing force and the weight point to balance the vibration causing force move away from each other, the moment that acts on the balancing shaft becomes large. According to the eighteenth aspect explained above, the vibration causing forces can be balanced by the corresponding elements, while the moment is suppressed in such a way that it is small enough.

According to the nineteenth aspect explained above, the connecting rod and the main shaft of the crankshaft can be connected by means of an eccentric shaft on the side of the crankshaft. According to the design, the crankshaft and connecting rod can relatively rotate with the bearing center on the side of the crankshaft located in the connecting rod as the center of rotation. Thus, the “connection point with the crankshaft” in the first aspect explained above can be formed in the center of the bearing on the side of the crankshaft. Further, according to the eccentric shaft on the crankshaft side, the center of the main shaft of the crankshaft can be set eccentric by a predetermined value from the center of the bearing on the side of the crankshaft, i.e. connection points with the crankshaft. Thus, according to the nineteenth aspect explained above, a “crankshaft coupling mechanism” can be specifically implemented that satisfies the function required by the first aspect explained above.

According to a twentieth embodiment, the connecting rod and the axial shaft of the balancing shaft can be connected by means of an eccentric shaft on the side of the balancing shaft. According to the design, the balancing shaft and the connecting rod can relatively rotate with the center of the bearing on the side of the balancing shaft located in the connecting rod as the center of rotation. Thus, the “connection point with the balancing shaft” in the first aspect explained above can be formed in the center of the bearing on the side of the balancing shaft. Additionally, according to the eccentric shaft on the side of the balancing shaft, the center of the axial shaft of the balancing shaft can be set eccentric by a fixed value from the center of the bearing on the side of the balancing shaft, i.e. connection points with balancing shaft. Thus, according to the twentieth aspect explained above, a “balancing shaft coupling mechanism" that specifically satisfies the function required by the first aspect explained above can be specifically implemented.

Brief Description of the Drawings

FIG. 1 is a view for explaining a configuration of a first embodiment of the present invention;

FIG. 2 is a diagram for explaining the relationship between the state of the balancing device shown in FIG. 1, and an angle θ of rotation of the crankshaft;

FIG. 3 is a diagram illustrating the relationship between the angle of rotation θ of the crankshaft and the angle α of rotation of the balancing shaft in the balancing device shown in FIG. one;

FIG. 4 is a diagram illustrating the relationship between the crank angle θ and the angular velocity dα / dθ of the balancing shaft in the balancing device shown in FIG. one;

FIG. 5 is a diagram illustrating the relationship between the angle θ of rotation of the crankshaft and the angular acceleration d2α / dθ2 of the balancing shaft in the balancing device shown in FIG. one;

FIG. 6 is a diagram for explaining the relationship between the operation of the piston and the angle θ of rotation of the crankshaft in the balancing device shown in FIG. one;

FIG. 7 illustrates the profiles of the inertia force generated by the piston of an internal combustion engine and causing vibration of the force generated by the balancing device shown in FIG. one;

FIG. 8 is a diagram showing a resultant vibration-inducing force obtained by combining a vibration-inducing force generated by a crankshaft and a vibration-inducing force generated by a balancing shaft, both of which are shown in FIG. 7;

FIG. 9 is a diagram showing the resultant force obtained by combining the inertia force generated by the piston and the resulting vibrational force, both of which are shown in FIG. 8;

FIG. 10 is a diagram for explaining a relationship between a configuration state of a modification example of the first embodiment of the present invention and a crank angle θ;

FIG. 11 illustrates profiles of the inertia force generated by the piston of an internal combustion engine and causing vibration of the force generated by the balancing device shown in FIG. 10;

FIG. 12 is a diagram showing the resultant vibration-inducing force obtained by combining the vibration-inducing force generated by the crankshaft and the vibration-inducing force generated by the balancing shaft, both of which are shown in FIG. eleven;

FIG. 13 is a diagram showing a resultant force obtained by combining an inertia force generated by a piston and a resultant vibration-inducing force, both of which are shown in FIG. 12;

FIG. 14 is a diagram for explaining the relationship between the state of the balancing device according to the second embodiment of the present invention and the angle θ of rotation of the crankshaft;

FIG. 15 is a diagram for explaining the relationship between the state of the balancing device according to the third embodiment of the present invention and the angle θ of rotation of the crankshaft;

FIG. 16 is a diagram showing a path of a connecting rod hinge during operation of the balancing device shown in FIG. fifteen;

FIG. 17 is a diagram for explaining conditions that must be satisfied by the guide section 82 illustrated in FIG. fifteen;

FIG. 18 illustrates an example of another configuration that can be used as a guide section in a third embodiment;

FIG. 19 is a diagram for explaining a configuration of a fourth embodiment of the present invention;

FIG. 20 is a diagram for explaining the operation of the balancing device shown in FIG. 19;

FIG. 21 illustrates the change in torque from the spring element that the balancing shaft receives from the spring element in each situation illustrated in FIG. twenty;

FIG. 22 illustrates a change in torque from a spring element that a balancing shaft receives from a spring element in a modification of a fourth embodiment of the present embodiment;

FIG. 23 is a diagram for explaining a configuration of a second modification of a fourth embodiment of the present invention;

FIG. 24 is a diagram for explaining a configuration of a third modification of a fourth embodiment of the present invention;

FIG. 25 is a diagram for explaining a configuration of a fifth embodiment of the present invention;

FIG. 26 is a diagram for explaining the operation of the balancing device shown in FIG. 25;

FIG. 27 illustrates the change in torque from the spring element that the balancing shaft receives from the spring element in each situation illustrated in FIG. 26;

FIG. 28 is a diagram for explaining a configuration of a sixth embodiment of the present invention;

FIG. 29 is a diagram for explaining an effect obtained by the construction of an offset crank located in the internal combustion engine shown in FIG. 28;

FIG. 30 is a diagram for explaining the relationship between the position of the piston in the internal combustion engine shown in FIG. 28, and the angle of rotation of the crankshaft;

FIG. 31A is a diagram for explaining a configuration of a sixth embodiment of the present invention;

FIG. 31B is a diagram for explaining a configuration of a first modification of a sixth embodiment of the present invention;

FIG. 31C is a diagram for explaining a configuration of a second modification of a sixth embodiment of the present invention;

FIG. 31D is a diagram for explaining a configuration of a third modification of a sixth embodiment of the present invention;

FIG. 32 is a diagram illustrating the relationship between the angle θ of rotation of the crankshaft and the angular velocity dα / dθ of the balancing shaft in the configuration shown in FIG. 31A using the bias coefficient h / r as a parameter;

FIG. 33 illustrates profiles of the inertia force generated by the piston of an internal combustion engine and the resulting vibrational force generated by the balancing device shown in FIG. 31A;

FIG. 34 is a diagram showing a result obtained by combining the resulting vibrational and inertial forces generated by the piston, both of which are shown in FIG. 33;

FIG. 35 is a diagram for explaining a configuration of a seventh embodiment of the present invention;

FIG. 36 is a diagram illustrating the relationship between the angle of rotation θ of the crankshaft and the angular velocity dα / dθ of the balancing shaft in the configuration shown in FIG. 35, using the displacement coefficient h / r as a parameter;

FIG. 37 illustrates profiles of the inertia force generated by the piston of an internal combustion engine and the resulting vibrational force generated by the balancing device shown in FIG. 35;

FIG. 38 is a diagram showing a result obtained by combining the resulting vibrational and inertia forces generated by a piston, both of which are shown in FIG. 37;

FIG. 39 is a diagram for explaining a configuration of an eighth embodiment of the present invention;

FIG. 40 is a diagram for explaining the principle of how the balancing device shown in FIG. 39, balances the forces causing vibration;

FIG. 41 is a diagram for explaining the position of the center of gravity of an eccentric load of a crankshaft located in the balancing device shown in FIG. 39;

FIG. 42 is a perspective view for explaining features of a crankshaft located in the balancing device shown in FIG. 39;

FIG. 43 is a perspective view for explaining the features of the balancing shaft located in the balancing device shown in FIG. 39;

FIG. 44 is a diagram for explaining the operation of the balancing device shown in FIG. 39;

FIG. 45 is a perspective view for explaining the features of another balancing shaft applicable for use in a balancing device according to an eighth embodiment of the present invention;

FIG. 46 is a diagram for explaining a configuration of a ninth embodiment of the present invention;

FIG. 47 is a diagram for explaining the principle of how the balancing device shown in FIG. 46, balances the forces causing vibration;

FIG. 48 is a diagram for explaining the position of the center of gravity of an eccentric load of a crankshaft located in the balancing device shown in FIG. 46;

FIG. 49 is a perspective view for explaining the features of the balancing shaft located in the balancing device shown in FIG. 46;

FIG. 50 is a diagram for explaining the operation of the balancing device shown in FIG. 46;

FIG. 51 is a perspective view for explaining the features of another balancing shaft applicable for use in a balancing device according to a ninth embodiment of the present invention;

FIG. 52 is an exploded perspective view of a balancing device according to a first embodiment of the present invention;

FIG. 53 is an exploded perspective view of the connecting rod shown in FIG. 52;

FIG. 54 is a side sectional view of a main section of a balancing device according to a tenth embodiment of the present invention;

FIG. 55 is a diagram for explaining a method for forming an eccentric shaft on the side of the crankshaft and an eccentric shaft on the side of the balancing shaft shown in FIG. 52, through general hole machining;

FIG. 56 is a diagram for explaining an operation of a balancing device according to a tenth embodiment of the present invention;

FIG. 57 is a diagram for explaining a first modification of a balancing device according to a tenth embodiment of the present invention;

FIG. 58 is a diagram for explaining a configuration of an eccentric shaft on a side of a balancing shaft located in a second modification of a tenth embodiment of the present invention;

FIG. 59 is a diagram for explaining a configuration of an eccentric shaft on a side of a crankshaft located in a third modification of a tenth embodiment of the present invention;

FIG. 60 is a diagram for explaining a configuration and operation of a fourth modification of a tenth embodiment of the present invention;

FIG. 61 is an exploded perspective view of a balancing device according to an eleventh embodiment of the present invention;

FIG. 62 is an exploded perspective view of the connecting rod shown in FIG. 61;

FIG. 63 is a side sectional view of a main section of a balancing device according to an eleventh embodiment of the present invention;

FIG. 64 is a diagram showing that an eccentric value occurring in an eccentric shaft on a balancing shaft side can be changed in an eleventh embodiment of the present invention;

FIG. 65 is a diagram for explaining an operation of a balancing device according to an eleventh embodiment of the present invention; and

FIG. 66 is a diagram for explaining a configuration and operation of a modification of an eleventh embodiment of the present invention.

DETAILED DESCRIPTION OF THE INVENTION

First Embodiment

Configuration of a First Embodiment

FIG. 1 is a view for explaining a configuration of a first embodiment of the present invention. The present embodiment includes an internal combustion engine 10. The internal combustion engine 10 has a piston 12. In the present embodiment, the internal combustion engine 10 is a single cylinder four stroke engine including only one piston 12.

The piston 12 is connected to the crankshaft 16 through a connecting rod 14 (hereinafter, referred to as "connecting rod 14"). The crankshaft 16 includes a crank pin 18 that connects to the connecting rod 14. The crank pin 18 is formed integrally with the crankshaft neck 22 (hereinafter referred to as the “crankshaft main shaft”) through the crank arm 20. The main shaft 22 of the crankshaft is rotatably held by a bearing located in the cylinder block.

The crankshaft 16 includes an eccentric load 24 of the crankshaft. An eccentric load 24 of the crankshaft is provided in such a way that its center of gravity is located on the almost opposite side of the crank pin 18 with the center of the main shaft 22 of the crankshaft between them. Additionally, the eccentric load 24 of the crankshaft is provided with a weight (mc) for balancing the weight of the connecting rod 14 and a weight for balancing the weight (mp / 2) corresponding to half the weight of the piston 12.

The crankshaft 16 performs one rotation at a time when the piston 12 performs one reciprocating movement between the top dead center and the bottom dead center. Hereinafter, the direction of reciprocating movement of the piston 12 is referred to as the "Y axis direction", and the direction orthogonal to the "Y axis direction" is called the "X axis direction".

In the present embodiment, the eccentric load 24 of the crankshaft is set so that the phase difference of the crankshaft angle of almost 180 ° (° CA) arises from the piston 12. Thus, the phase of the eccentric load 24 of the crankshaft is set so that the following two conditions:

(1) when the piston 12 is located at top dead center, the center of gravity of the eccentric load 24 of the crankshaft is located practically on the moving end on the side of the bottom dead center in the direction along the Y axis;

(2) when the piston 12 is located at bottom dead center, the center of gravity of the eccentric load 24 of the crankshaft is located practically at the movable end on the side of top dead center in the Y axis direction.

The internal combustion engine 10 includes a balancing device 30. The crankshaft 16 is a component of the balancing device 30. The crankshaft 16 contains a connection point 32 with the crankshaft at a position deviating from the center of the main shaft 22 of the crankshaft. More specifically, the connection point 32 with the crankshaft is provided on the opposite side of the eccentric load 24 of the crankshaft with the center of the main shaft 22 of the crankshaft between them.

The connecting rod 36 is connected to the crankshaft connection point 32 through the crankshaft connection mechanism 34. The connecting rod 36 is rotatably held by a crankshaft coupling mechanism 34. Therefore, the crankshaft 16 and the connecting rod 36 can relatively rotate in a plane parallel to the surface of rotation of the crankshaft 16, with a connection point 32 with the crankshaft as the center of rotation.

The other end of the connecting rod 36 is connected to the balancing shaft 40 at a point 38 of the connection with the balancing shaft. The balancing shaft 40 includes a mechanism 42 for connecting to the balancing shaft at a point 38 for connecting to the balancing shaft. The connecting rod 36 is rotatably held by means of a mechanism 42 for connecting to the balancing shaft. Therefore, the connecting rod 36 and the balancing shaft 40 can relatively rotate with the point 38 of the connection with the balancing shaft as the center of rotation.

The balancing shaft 40 includes an axial shaft 44 of the balancing shaft in a position deviated from the point 38 of the connection with the balancing shaft. The axial shaft 44 of the balancing shaft is provided parallel to the main shaft 22 of the crankshaft and is rotatably held by a bearing engaged by the cylinder block. Therefore, the balancing shaft 40 can rotate in a plane that is parallel to the surface of rotation of the crankshaft 16, with the axial shaft 44 of the balancing shaft as the rotational shaft.

The balancing shaft 40 includes a mechanism 46 for adjusting the connection points with the balancing shaft, which holds the mechanism 42 for connecting to the balancing shaft. The mechanism 46 for adjusting the points of connection to the balancing shaft is provided in such a way that the position of the mechanism 42 for connecting to the balancing shaft on the balancing shaft 40, i.e. the position of the connection point 38 with the balancing shaft is passively adjusted to the correct position. By the function of the balancing shaft connection point adjustment mechanism 46, the balancing shaft connection point 38 can be shifted in a fixed range in the direction of the rotation radius of the balancing shaft 40, which passes through the axial shaft 44 of the balancing shaft.

The balancing shaft 40 further comprises an eccentric load 48 of the balancing shaft. The eccentric load 48 of the balancing shaft is provided so that its center of gravity is located on the opposite side of the connection point 38 with the balancing shaft with the center of the axial shaft 44 of the balancing shaft between them. Additionally, the eccentric load 48 of the balancing shaft is provided with a weight to balance the weight (mp / 2), almost corresponding to half the weight of the piston 12.

In the present embodiment, while the crankshaft 16 performs one rotation, the balancing shaft 40 performs one rotation in the opposite direction with respect to the rotation direction of the crankshaft 16. Here, as illustrated in FIG. 1, the eccentric load 48 of the balancing shaft is provided so that it is in phase synchronized with the eccentric load 24 of the crankshaft. Thus, the phase of the eccentric load 48 of the balancing shaft is shifted from the phase of the piston 12 by almost 180 ° CA, similarly to the phase of the eccentric load 24 of the crankshaft. Therefore, the two conditions described as follows are also established between the phase of the eccentric load 48 of the balancing shaft and the phase of the piston 12:

(1) when the piston 12 is located at top dead center, the center of gravity of the eccentric load 48 of the balancing shaft is located practically at the movable end on the side of the bottom dead center in the direction along the Y axis;

(2) when the piston 12 is located at bottom dead center, the center of gravity of the eccentric load 48 of the balancing shaft is located practically on the movable end on the side of top dead center in the direction along the Y axis.

The connecting rod 36 includes a hinge 50 at its midpoint. The circular sliding portion 52 is inserted into the hinge 50. The cylinder block includes a guide section 54 that controls the movement of the sliding part 52. The guide section 54 has a sliding space having a longitudinal direction in the direction of the center line 56 "crankshaft - balancing shaft", which passes through the center the main shaft 22 of the crankshaft and the center of the axial shaft 44 of the balancing shaft. The sliding portion 52 may move along the inner wall of the sliding space. As a result, the movement of the hinge 50 is limited to rectilinear movement on the center line 56 "crankshaft - balancing shaft".

Explanation of the basic mode of operation of the balancing device of the first embodiment of the invention

FIG. 2 is a diagram for explaining the operation of the balancing device 30 illustrated in FIG. 1. In FIG. 2, the horizontal axis represents the angle θ of rotation of the crankshaft (° CA). Hereinafter, the crank angle θ of 0 (° CA) or 360 (° CA) corresponds to top dead center, and the crank angle θ of 180 (° CA) corresponds to bottom dead center. Additionally, the direction of the center line 56 "crankshaft - balancing shaft" is specified as the "y direction", and the direction orthogonal to the "direction along the y axis" is set as the "direction along the x axis." It should be noted that, in FIG. 2, for convenience of explanation, the y-axis direction is shown in such a way that it corresponds to the y-axis direction (piston movement direction 12).

In FIG. 2, in the state θ = 0 (° CA), the angle of rotation α of the balancing shaft 40 (hereinafter referred to as the "angle α of rotation of the balancing shaft") is also 0 (degrees). At this time, the center of gravity of the eccentric load 24 of the crankshaft and the center of gravity of the eccentric load 48 of the balancing shaft are at the movable ends on the bottom dead center side.

In FIG. 2, the crankshaft 16 is supposedly rotated in a clockwise direction. When the crankshaft 16 rotates from the state θ = 0 (° CA), the connection point 32 with the crankshaft shifts in the direction (positive direction along the x axis, negative direction along the y axis). At this time, the x coordinate of the hinge 50 is always maintained at the starting point by the guide section 54. Therefore, the connection point 38 with the balancing shaft is shifted in the direction (negative direction along the x axis, negative direction along the y axis). As a result, the balancing shaft 40 rotates in the opposite direction relative to the direction of rotation of the crankshaft 16.

Until θ reaches 90 (° CA), the connection point 32 with the crankshaft continues to shift in the direction (positive direction along the x axis, negative direction along the y axis). When the x coordinate of point 38 of the connection to the balancing shaft changes by a distance identical to the distance of the x coordinate of point 32 of the connection to the crankshaft at that time, i.e. when the connecting rod 36 maintains a vertical state in FIG. 2, the y coordinate of the balancing shaft connection point 38 is shifted by an equivalent distance with the y coordinate of the crankshaft connection point 32. However, in the balancing device 30, the connection point 32 with the crankshaft and the connection point 38 with the balancing shaft are moved apart from each other in the x-axis direction until θ reaches 90 (° CA). In order to compensate for the distance in the x-axis direction, the y coordinate of the point 38 of the connection to the balancing shaft cannot deter a significant offset from the y coordinate of the point 32 of the connection to the crankshaft. As a result, at the stage where the crankshaft 16 rotates 90 (° CA), the angle α of rotation of the balancing shaft exceeds 90 (degrees).

As long as the angle of rotation of the crankshaft does not exceed 90 (° CA), to reach 180 (° CA), the point 32 of the connection with the crankshaft changes in the direction (negative direction along the x axis, negative direction along the y axis). At this time, the connection point 38 with the balancing shaft is shifted in the direction (positive direction along the x axis, negative direction along the y axis). Here, the connection point 32 with the crankshaft and the connection point 38 with the balancing shaft approach each other in the x axis direction with increasing angle θ of rotation of the crankshaft. Therefore, the change in the y coordinate of the connection point 38 with the balancing shaft is less than the change in the y coordinate of the connection point 32 of the crankshaft. Then, at the stage at which the angle of rotation of the crankshaft reaches 180 [° CA], the angle α of rotation of the balancing shaft likewise reaches 180 (degrees).

For the above reason, the balancing shaft 40 rotates at a higher rotation speed than the crankshaft 16 in the process of changing the angle of rotation θ of the crankshaft from 0 (° CA) to 90 (° CA). Then, the rotation speed of the balancing shaft 40 is lower than the rotation speed of the crankshaft 16 while changing the angle of rotation θ of the crankshaft from 90 (° CA) to 180 (° CA). This change in rotational speed also occurs when the crank angle θ is changed from 180 (° CA) to 360 (° CA) by a similar mechanism.

As described above, the balancing device 30 in the present embodiment has the following characteristics:

(1) the crankshaft 16 and the balancing shaft 40 rotate in opposite directions in an identical cycle;

(2) when the crankshaft 16 rotates at an equal rotation speed, rotation at an unequal rotation speed occurs in the balancing shaft 40. In this case, the rotation speed of the balancing shaft 40 is higher than the rotation speed of the crankshaft 16, when the rotation angle θ of the crankshaft belongs to the range 0 ( ° CA) -90 (° CA) and a range of 270 (° CA) -360 (° CA). Additionally, when the angle of rotation of the crankshaft belongs to the range 90 (° CA) -270 (° CA), the rotation speed of the balancing shaft 40 is lower than the rotation speed of the crankshaft 16;

(3) in the state of the crank angle θ = 0 (° CA), i.e. in a state in which the piston 12 is located at top dead center, the center of gravity of the crankshaft 16 and the center of gravity of the balancing shaft 40 are at the movable end at bottom dead center. Additionally, in the state of the crank angle θ = 180 (° CA), i.e. in a state in which the piston 12 is located at bottom dead center, the center of gravity of the crankshaft 16 and the center of gravity of the balancing shaft 40 are on the movable end on the top dead center side.

During operation of the internal combustion engine 10, a reciprocating movement occurs in the piston 12, and rotational movements occur in the crankshaft 16 and the balancing shaft 40. At this time, a composite movement of the reciprocating movement and the rotational movement occurs in the connecting rod 14. The main the weight of the connecting rod 14 exists in the part that rotates with the crank pin 18. Therefore, in the weight (mc + mp / 2) of the crankshaft 16, (mc) is balanced by the rotational part of the connecting rod 14. Accordingly, during operation of the internal combustion engine 10, it can be considered that the following movement occurs in the internal part of the internal combustion engine.

(1) Reciprocating movement in the Y direction of the load (mp) due to movement of the piston 12

(2) Normal rotational movement of the eccentric load (mp / 2) due to rotation of the crankshaft 16

(3) Reverse rotational movement of the eccentric load (mp / 2) due to rotation of the balancing shaft 40

The reciprocating movement in the direction along the Y axis of the load (mp) generates the inertia force in the direction along the Y axis. The inertia force changes in absolute value synchronously with the movement of the piston 12, reaches an almost negative maximum value at top dead center and reaches an almost positive maximum value at bottom dead center.

The normal rotational movement of the eccentric load (mp / 2) and the reverse rotational movement of the eccentric load (mp / 2) form vibration-causing forces directed outward from the respective radii of rotation. The components on the X axis of the vibration-causing forces are balanced by each other, and the components on the Y axis are combined. The combined components along the Y axis of the vibration-inducing force balance the inertia that accompanies the movement of the piston. Therefore, according to the balancing device 30 in the present embodiment, vibration in the operation of the internal combustion engine 10 can be suppressed so that it is small enough.

Detailed explanation of the operation of the balancing device of the first embodiment of the invention

FIG. 3 illustrates the relationship that is established between the angle of rotation θ of the crankshaft and the angle α of rotation of the balancing shaft in the present embodiment. FIG. 3 shows a state in which the above rotation at an unequal rotation speed occurs in the balancing shaft 40 with the rotation of the crankshaft 16.

FIG. 4 is a diagram in which the angle of rotation α of the balancing shaft illustrated in FIG. 3 is replaced by the angular velocity dα / dθ of the balancing shaft. When the balancing shaft 40 rotates at an equal speed of rotation in the opposite direction relative to the crankshaft 16, the angular velocity dα / dθ of the balancing shaft is always -1. In contrast, in the present embodiment, the angular velocity dα / dθ of the balancing shaft has a value with a large absolute value in the vicinity of the top dead center due to rotation at an unequal rotation speed of the balancing shaft 40. Additionally, in the vicinity of the bottom dead center, the angular velocity dα / dθ of the balancing shaft matters with a small absolute value.

FIG. 5 is a diagram in which the angular velocity dα / dθ of the balancing shaft, illustrated in FIG. 4 is further replaced by angular acceleration d2α / dθ2 of the balancing shaft. When the rotation speed of the balancing shaft 40 is equal to the rotation speed of the crankshaft 16, the angular acceleration d2α / dθ2 of the balancing shaft is always zero. In contrast, in the present embodiment, the angular acceleration d2α / dθ2 of the balancing shaft has a value with a large absolute value in the vicinity of the middle of the top dead center and bottom dead center due to rotation at an unequal rotation speed of the balancing shaft 40.

A centrifugal force, which is proportional to the square of the angular velocity, acts on the balancing shaft 40. Additionally, when angular acceleration occurs in the balancing shaft 40, the angular acceleration reaction force acts on the balancing shaft 40. The balancing shaft 40 generates a vibration-causing force corresponding to the composite value from the above centrifugal force and reaction force. Therefore, the vibration causing force that is generated by the balancing shaft 40 in the present embodiment is distorted compared to the vibration causing force that is generated by the crankshaft 16.

FIG. 6 is a diagram for explaining the movement of the piston 12 with respect to the change in the angle of rotation θ of the crankshaft. In the internal combustion engine 10, the piston 12 is connected to the main shaft 22 of the crankshaft via the crankshaft 16 and the connecting rod 14. Here, the length of the connecting rod is set to 1c and the radius of the crank is set to rc. Additionally, the ratio of the length of the connecting rod and the radius of the crank, i.e. connecting rod ratio expressed by 1c / rc.

As illustrated in FIG. 6, in the process of changing the angle θ of rotation of the crankshaft from 0 (° CA) to 90 (° CA), the coordinates of the crank pin 18 are shifted in the direction (positive direction along the x axis, negative direction along the y axis). Since the X coordinate of the piston 12 is fixed, the crank pin 18 moves away from the piston 12 in the X direction in this process. In order to compensate for the distance in the X-axis direction, the Y coordinate of the piston 12 cannot be held back from a more significant displacement compared to the Y coordinate of the crank pin 18. Therefore, the stroke PS90, which occurs in the piston 12 during a change in the crank angle θ from 0 (° CA) to 90 (° CA), exceeds the amount of displacement in the direction along the Y axis that occurs in the crank pin 18 during this time .

In the process of changing the angle θ of rotation of the crankshaft to 180 (° CA) from 90 (° CA), the coordinates of the crank pin 18 are shifted in the direction (negative direction along the x axis, negative direction along the y axis). In this process, the crank pin 18 approaches the piston 12 in the X-axis direction. Since both the crank pin 18 and the piston 12 approach each other in the X-axis direction, the amount of displacement in the Y-axis direction of the piston 12 becomes smaller compared to with the amount of displacement in the direction along the Y axis of the crank pin 18. Therefore, the amount (PS180-PS90) of the piston displacement during the change of the angle θ of rotation of the crankshaft from 90 (° CA) to 180 (° CA) becomes less than the above PS90. A similar change in displacement occurs when the piston 12 moves from the bottom dead center side to the top dead center side. For the above reason, in the case of rotation at an equal rotational speed of the crankshaft 16, the speed of the displacement piston 12 becomes relatively high in the vicinity of the top dead center and becomes relatively low in the vicinity of the bottom dead center.

FIG. 7 illustrates the profiles of the inertia force generated by the piston 12 and causing the vibration of the force generated by the balancing device 30. The waveforms 60, 62 and 64, respectively, are force profiles described as follows:

Waveform 60: component along the Y axis of the vibration-inducing force generated by the eccentric load (mp / 2) of the crankshaft 16

Waveform 62: Y-axis component of the vibration-inducing force generated by the eccentric load (mp / 2) of the balancing shaft 40

Waveform 64: Inertia generated by piston 12

The crankshaft 16 causes the eccentric load (mp / 2) to rotate at an equal speed of rotation. Therefore, the waveform 60 corresponding to the vibrating force of the crankshaft 16 is a sine wave with virtually no distortion.

The balancing shaft 40 causes the eccentric weight (mp / 2) to rotate at an unequal speed of rotation quickly in the process from top dead center to mid point and slowly in the process from mid point to bottom dead point. Therefore, the waveform 62 corresponding to the balancing shaft 40 is in the form of a distorted sine wave having shoulders in the vicinity of 90 (° CA) and in the vicinity of 270 (° CA).

The piston 12 generates an inertia force corresponding to its displacement speed. The speed of the displacement piston 12 becomes high in the vicinity of the top dead center and becomes low in the vicinity of the bottom dead center, as described above. Therefore, the waveform 64 corresponding to the piston 12 is in the form of a distorted sine wave that has peaks at 0 (° CA) or 360 (° CA) on the top dead center side, but does not have a peak in the vicinity of 180 (° CA) on the side bottom dead center.

FIG. 8 is a diagram showing an inertia force of a piston 12 with a vibration-inducing force obtained by combining a component along the Y-axis of the vibration-generating force generated by the crankshaft 16 and a component along the Y-axis of the vibration-inducing force generated by the balancing shaft 40. Three kinds of waveforms illustrated in FIG. 8, respectively, are as described below.

Waveform 60 * 2: corresponds to twice the vibrational force of the waveform 60 illustrated in FIG. 7

Waveform 60 + 62: corresponds to a composite object of waveform 60 and waveform 62 illustrated in FIG. 7

Waveform 64: is identical to waveform 64 illustrated in FIG. 7

In the present embodiment, the rotation of the crankshaft 16 is transmitted to the balancing shaft 40 through the connecting rod 36. In contrast, the rotation of the crankshaft 16 can also be transmitted to the balancing shaft 40 by using a conventional gear train, for example, in the form of a perfect circle. In this case, the balancing shaft 40 rotates at an equal rotational speed relative to the rotational speed of the crankshaft 16, and the y-axis component of the vibrating force of the balancing shaft 40 changes along a sinusoidal wave without distortion, identical to the waveform 60. Accordingly, in this case, the resulting force of the components along the Y axis of the vibration-causing force generated by the crankshaft 16 and the balancing shaft 40 should correspond to the resulting force, which is obtained by doubling the vibration-generating force of the waveform 60. Thus, the resulting force must correspond to the waveform 60 * 2 illustrated in FIG. 8.

Waveform 60 does not include distortion. Therefore, even when the waveform 60 is combined with the waveform 60, the resulting waveform 60 * 2 does not approach the waveform of the inertia of the piston 12 so much. In contrast, the waveform 60 + 62 generated by combining the waveform 60 with the waveform 62 is relatively flat in the vicinity of the bottom dead center and has a peak-like increase in the vicinity of the top dead center. The waveform 60 + 62 is much closer to the symmetrical waveform 64, compared to the waveform 60 * 2.

FIG. 9 illustrates waveforms generated by further combining the waveform 60 * 2 and the waveform 60 + 62 with the waveform 64 illustrated in FIG. 8, respectively. The semantic values of the corresponding waveforms are described as follows.

Waveform 60 * 2 + 64: unbalanced force remaining in the internal combustion engine 10 when the balancing shaft 40 is controlled by a gear

Waveform 60 + 62 + 64: unbalanced force remaining in the internal combustion engine 10 in the present embodiment

As shown by the waveform 60 + 62 + 64, the unbalanced force remaining in the internal combustion engine 10 in the present embodiment is sufficiently small in the entire region of the crankshaft angle θ. The unbalanced force is quite small compared to the unbalanced force in the case of rotation of the balancing shaft 40 at an equal speed of rotation with a gear transmission (signal shape 60 * 2 + 64).

The connecting rod 36 used in the present embodiment can be formed in such a way that it is much lighter and more compact in comparison with the gear transmission. Therefore, the configuration of the present embodiment is more advantageous for reducing the size and weight of the internal combustion engine 10 and can provide more superior smoothness to the internal combustion engine 10 compared to the case of the balancing shaft 40 being rotated by using a gear train.

An example of a modification of the first embodiment of the invention

FIG. 10 is a diagram for explaining a configuration of an example modification of a first embodiment of the present invention. In the aforementioned first embodiment, the connection point 32 with the crankshaft is provided on the opposite side of the eccentric load 24 of the crankshaft with the center of the main shaft 22 of the crankshaft between them, and the connection point 38 with the balancing shaft is provided on the opposite side of the eccentric load 48 of the balancing shaft with the center of the axial the shaft 44 of the balancing shaft between them (see Fig. 1). However, the configuration of the present invention is not limited to this. Thus, as illustrated in FIG. 10, a connection point 32 with a crankshaft can be provided on a side identical to the side of the eccentric load 24 of the crankshaft relative to the center of the main shaft 22 of the crankshaft, and a connection point 38 with the balancing shaft can be provided on the side identical to the side of the eccentric load 48 of the balancer shaft shaft 44 of the balancing shaft.

FIG. 11, 12 and 13 illustrate waveforms of inertial forces and vibration-causing forces that are generated when operating in the above modification example. The waveforms illustrated in the respective drawings are assigned common reference numbers with reference numbers with the numbers shown in FIG. 7, 8, and 9. As illustrated in FIG. 13, according to the configuration of the present modification example, the unbalanced force (waveform 60 + 62 + 64) remaining in the internal combustion engine 10 can be set smaller compared with the case of the first embodiment. It should be noted that the positions of the connection point 32 with the crankshaft and the connection points 38 with the balancing shaft are not limited to the positions illustrated in FIG. 10, and can be appropriately set in accordance with the waveform of the component along the Y axis of the vibration causing force to be generated in the balancing device 30.

Further, in the above first embodiment, the internal combustion engine 10 is described as a single cylinder engine, but the configuration of the present invention is not limited thereto. For example, in a four-stroke two-cylinder engine, the pistons of the two cylinders perform reciprocating movements in the identical phase. The present invention can be used to balance inertia generated by two pistons.

Additionally, in the above first embodiment, the phase of the center of gravity of the crankshaft 16 and the phase of the center of gravity of the balancing shaft 40 are set so that they correspond to each other at top dead center and bottom dead center, but the phases of both of them can be set in such a way, if necessary that they are different from each other in order to obtain the required force causing vibration. The phases of both of them, in particular, may differ according to need within a range of 45 (° CA), within a range of 30 (° CA), within a range of 15 (° CA), or within a range of 5 ( ° CA).

It should be noted that the above three modifications can be used not only as a modification of the first embodiment, but also be used as modifications of all other embodiments, which are described below.

Further, in the above first embodiment, the sliding portion 52 which is held by the guide section 54 is circular, but the configuration of the present invention is not limited thereto. The sliding portion 52 may be anything that may cause the hinge 50 to perform a rectilinear movement along the guide section 54, and its shape may be a square shape or a rectangular shape in which corners are excluded.

Additionally, in the above first embodiment, the hinge 50, which is held by the guide section 54, is provided at the midpoint of the connecting rod 36, but the configuration of the present invention is not limited to this. Thus, the hinge 50 of the connecting rod 36 can be provided at an arbitrary point on the connecting rod 36 within the range in which the balancing device 30 is operated.

Further, in the above first embodiment, the balancing shaft connection point adjusting mechanism 46 is provided in the balancing shaft 40, but the configuration of the present invention is not limited thereto. Thus, a mechanism for adjusting the connection points can be provided in each of the balancing shaft 40 and the crankshaft 16, or in the crankshaft 16, so that the balancing device 30 can operate.

Second Embodiment

Next, a second embodiment of the present invention will be described with reference to FIG. 14. FIG. 14 is a diagram for explaining the operation of the balancing device 66 in the present embodiment. The configuration of the present embodiment can be realized by mounting the balancing device 66 illustrated in FIG. 14, on the internal combustion engine 10 illustrated in FIG. 1, instead of the balancing device 30 illustrated in FIG. 2. Hereinafter, in FIG. 14, common elements with elements illustrated in FIG. 2, common references with numbers are assigned, and their explanation is omitted or simplified.

The balancing device 66 illustrated in FIG. 14 includes a guide section 68 to adjust the path of the connecting rod 36. The guide section 68 has a groove 70 substantially equal to the width of the connecting rod 36. The connecting rod 36 can slide in the groove 70 in a state in which the connecting rod 36 is located in the groove 70.

The guide section 68 includes a rotational shaft 72 in a portion in which the connecting rod 36 is superimposed in FIG. 14 (initially located on the rear side of the connecting rod 36 and cannot be visually recognized). The rotary shaft 72 is parallel with the main shaft 22 of the crankshaft and the axial shaft 44 of the balancing shaft and is provided in a position superimposed on the center line 56 of the "crankshaft - balancing shaft" connecting the center of the main shaft 22 of the crankshaft and the center of the axial shaft 44 of the balancing shaft. Therefore, the guide section 68 can rotate in the moving plane of the connecting rod 36 while holding the connecting rod 36.

In FIG. 14, a black circle point 74 drawn on the connecting rod 36 expresses the midpoint of the connecting rod 36. In the above first embodiment, the movement of the connecting rod 36 is controlled so that the midpoint (hinge 50) of the connecting rod 36 moves on the center line 56 "crankshaft - balancing shaft "(see Fig. 2). In this case, the connecting rod 36 always transfers the displacement that occurs at the point 32 of the connection with the crankshaft to the point 38 of the connection with the balancing shaft with a 1: 1 lever.

In the present embodiment, as illustrated in FIG. 14, in the situation of the crank angle θ = 0 (° CA), the midpoint 74 of the connecting rod 36 deviates from the guide section 68 to the upstream side, i.e. to the side of the connection point 38 with the balancing shaft. In this case, the displacement that occurs with the connection point 32 with the crankshaft is transmitted to the connection point 38 with the balancing shaft when amplified with a lever greater than one. Then, when the angle of rotation of the crankshaft exceeds 90 (° CA), the midpoint 74 of the connecting rod 36 is in a state deflected toward the lower side of the guide section 68, i.e. to the side of the point 32 of the connection with the crankshaft. At this stage, the displacement that occurs on the side of the connection point 32 with the crankshaft is transmitted to the connection point 38 with the balancing shaft when decreasing with the lever less than unity.

Thus, according to the configuration of the present embodiment, the displacement that occurs at the connection point 32 of the crankshaft can be transmitted to the connection point 38 to the balancing shaft properly with various levers. More specifically, in a region in the vicinity of the top dead center where the crank angle θ is about 0 (° CA), the balancing shaft 40 can rotate at a higher rotation speed than the case of the first embodiment relative to the rotation of the crank angle θ shaft. In the region in the vicinity of the bottom dead center where the crank shaft angle θ is about 180 (° CA), the balancing shaft 40 can rotate more slowly compared to the case of the first embodiment with respect to the rotation of the crank shaft angle θ.

When the profile of the rotation speed of the balancing shaft 40 changes, the profile of the vibration causing force generated by the balancing shaft 40 is also different. Therefore, according to the balancing device 66 in the present embodiment, a vibration causing force profile different from the case of the first embodiment can be formed.

The inertia force generated by the piston 12 of the internal combustion engine 10 exhibits different profiles in accordance with different design values. According to the configuration of the present embodiment, excellent smoothness of operation can be ensured for the internal combustion engine 10 if the profile of the inertia force generated by the piston 12 is substantially similar to the profile of the vibration causing force generated by the balancing device 66 illustrated in FIG. fourteen.

Third Embodiment

The following describes a third embodiment of the present invention with reference to FIG. 15-18. FIG. 15 is a diagram for explaining the operation of the balancing device 78 in the present embodiment. The configuration of the present embodiment can be realized by mounting the balancing device 78 illustrated in FIG. 15, on the internal combustion engine 10 illustrated in FIG. 1, instead of the balancing device 30 illustrated in FIG. 2. The balancing device 78 is similar to the balancing device 30 illustrated in FIG. 2, with the exception of three aspects described as follows. It should be noted that in FIG. 15, to common elements with the elements illustrated in FIG. 2, common references with numbers are assigned, and their explanation is omitted or simplified.

Difference 1

The balancing device 78 of the present embodiment includes a balancing shaft 80. The mechanism 42 for connecting to the balancing shaft is directly connected to the balancing shaft 80 without using the mechanism 46 for adjusting the points of connection with the balancing shaft (see Fig. 2).

Difference 2

The balancing device 78 of the present embodiment satisfies the following conditions:

(1) The distance L _CB between the center of the main shaft 22 of the crankshaft and the center of the axial shaft 44 of the balancing shaft is equal to the distance l _cb between the connection point 32 to the crankshaft and the connection point 38 to the balancing shaft.

(2) The radius r1 of rotation of the point 32 of the connection with the crankshaft is equal to the radius r2 of rotation of the point 38 of the connection with the balancing shaft.

Difference 3

The balancing device 78 of the present embodiment includes a guide section 82. In FIG. 15, a figure-eight dashed line illustrated in the guide section 82 shows the path 84 of the hinge 50. The width of the guide section 82 is greater than the diameter of the sliding part 52, so that the hinge 50 can move along the path 84.

In FIG. 15, in the condition of the angle of rotation θ of the crankshaft at 0 (° CA), the center line of the connecting rod 36 is superimposed on the line connecting the center of the axial shaft 44 of the balancing shaft and the connection point 38 to the balancing shaft. Later in this document, this point is called the "first point of change." At the first point of change, the axial force that is provided to the connection point 38 with the balancing shaft from the connecting rod 36 does not provide torque to the balancing shaft 80. Therefore, the balancing shaft 80 is in a state of rotation, both in the direction of normal rotation and in the direction of reverse rotation at the first point of change.

When the crankshaft 16 normally rotates from the first point of change, a slight displacement in the direction (positive direction along the x axis, negative direction along the y axis) occurs at the point 32 of connection with the crankshaft. The distance between the connection point 32 with the crankshaft and the connection point 38 with the balancing shaft is always constant, and therefore, if the above-described displacement occurs at the connection point 32 with the crankshaft, the connection point 38 with the balancing shaft cannot be prevented from shifting in any of the directions to compensate for the offset.

If the connecting rod 36 can move freely, the connection point 38 with the balancing shaft can be displaced both in the positive direction along the x axis and in the negative position along the x axis, i.e. both in the direction of normal rotation and in the direction of reverse rotation. When the connection point 38 with the balancing shaft is displaced in the normal rotation direction with the connection point 32 of the crankshaft, the hinge 50 in the connecting rod 36, which is located at the midpoint between the connection point 32 with the crankshaft and the connection point 38 with the balancing shaft, is necessarily shifted in the direction normal rotation along an ideal circular path. On the other hand, when the connection point 38 with the balancing shaft is displaced in the opposite direction relative to the connection point 32 with the crankshaft, a small x-axis offset and a large y-axis offset occur in the hinge 50, which is located at the midpoint between the crankshaft connection 32 shaft and point 38 connection with the balancing shaft. In this case, the hinge 50 moves along the figure-eight path 84.

FIG. 16 is a diagram showing a path 84 (hereinafter referred to as a "reverse rotation path") in the form of a figure eight described above, different from a path 86 (hereinafter referred to as a "normal rotation path") of the hinge 50 according to the normal the rotation of the point 38 of the connection with the balancing shaft. It should be noted that the slide path 88 illustrated by assigning the reference number 88 in FIG. 16 expresses the shape that the outer wall of the sliding portion 52 adheres to when the hinge 50 moves along the reverse rotation path 84.

FIG. 17 is a diagram for explaining the details of the guide section 82 illustrated in FIG. 15. As illustrated in FIG. 17, the guide section 82 includes a guide 90 on the side of the balancing shaft at one end. A guide 90 on the side of the balancing shaft is provided at the point at which the sliding portion 52 is located at the first change point and has a shape identical to that of the upper end portion of the sliding path 88 illustrated in FIG. 16. The guide 90 on the side of the balancing shaft can prevent the sliding portion 52 from moving along the normal rotation path 86 and cause the sliding portion 52 to move along the reverse rotation path 84 at the first change point.

For the reason described above, in the balancing device 78 of the present embodiment (see FIG. 15), during a slight rotation of the crankshaft 16 from the first change point, the sliding portion 52 can always move along the reverse rotation path 84. As a result, the direction of displacement of the connection point 38 with the balancing shaft necessarily becomes the direction of reverse rotation, i.e. negative direction along the x axis. When the connection point 38 with the balancing shaft even moves slightly in the negative direction along the x axis from the first change point, the axial force generated by the connecting rod 36 provides torque to the balancing shaft 40. After that, until the crank shaft angle θ of rotation reaches 180 (° CA), balancing shaft 80 receives torque and maintains stable reverse rotation.

When the angle of rotation of the crankshaft reaches 180 (° CA), a state in which the line connecting the center of the axial shaft 44 of the balancing shaft and the connection point 38 to the balancing shaft is superimposed on the center line of the connecting rod 36 is formed again. Later in this document, the point is called the "second point of change." At the second change point, the balancing shaft 80 is in a state in which the balancing shaft 80 can rotate both in the normal rotation direction and in the reverse rotation direction, similarly to the first change point.

As illustrated in FIG. 17, the guide section 82 in the present embodiment includes a guide 92 on the side of the crankshaft on the opposite side of the guide 90 on the side of the balancing shaft. A guide 92 on the side of the crankshaft is provided at the point at which the sliding portion 52 is located at the second change point and has a shape identical to that of the lower end portion of the sliding path 88 illustrated in FIG. 16. A guide 92 on the side of the crankshaft can prevent the sliding portion 52 from moving along the normal rotation path 86 and cause the sliding portion 52 to move along the reverse rotation path 84 at the second change point.

Due to the adjustment by the guide 92 on the side of the crankshaft, the sliding part 52 always moves along the reverse rotation path 84 when the crankshaft 16 rotates from the second change point. When a slight bias occurs, the torque in the reverse direction of rotation acts on the balancing shaft 80 stably until the crankshaft 16 returns to the first point of change. The above operation is repeated, due to which, in the present embodiment, the balancing shaft 80 is also provided with the ability to perform rotation at an unequal speed of rotation in the opposite direction relative to the direction of rotation of the crankshaft 16.

As illustrated in FIG. 16, while the hinge 50 moves along the reverse rotation path 84, the outer wall of the sliding portion 52 adheres to the sliding path 88 having a tapered portion in the central portion. Therefore, the inner wall of the guide section 82 can be formed by a set of curved lines, similar to the sliding path 88. However, the function required by the guide section 82 in the present embodiment is only to control the movement of the sliding part 52 at the first change point and the second change point.

In other words, if the movement of the sliding part 52 can be properly controlled at the first change point and the second change point in the present embodiment, in the process between the first change point and the second change point, the balancing shaft 80 can force the reverse rotation stably by the axial force of the connecting rod 36 Therefore, the guide section 82 does not have to be in contact with the sliding portion 52 between the first change point and the second change point.

In FIG. 17, the guide section 82 includes straight side walls 94 and 96 between the guide 90 on the side of the balancing shaft and the guide 92 on the side of the crankshaft. Since these side walls 94 and 96 are rectilinear and, therefore, are easily formed compared to the side walls in curved forms of the sliding path 88 (see FIG. 16). Additionally, these side walls 94 and 96 do not interfere with the sliding portion 52, which moves along the sliding path 88. Therefore, according to the guide section 82, as illustrated in FIG. 17, the desired function can be realized while simplifying the manufacturing process.

As described above, by the configuration of the present embodiment, the balancing shaft 80 can also be forced to rotate at an unequal rotation speed in the opposite direction with respect to the direction of rotation of the crankshaft 16 by means of a structure using a connecting rod 36. Therefore, according to the configuration of the present embodiment, can also be realized compact internal combustion engine 10, which has excellent smooth operation, as in the case of the first sludge second embodiment.

It should be noted that in the above third embodiment, the hinge 50 and the sliding part 52 of the connecting rod 36 correspond to the “restrictive part” in the fourth embodiment of the invention.

An example of a modification of a third embodiment of the invention

FIG. 18 illustrates an example of another configuration that can be used as the guide section 82 in the third embodiment. In the above third embodiment, straight side walls 94 and 96 are provided between the guide 90 on the balancing shaft side and the guide 92 on the crankshaft side. In the example illustrated in FIG. 18, these side walls 94 and 96 are lowered. The guide section 82 in the third embodiment may be implemented by such a simple configuration.

Fourth Embodiment

Next, a fourth embodiment of the present invention is described with reference to FIG. 19-24. FIG. 19 is a diagram for explaining a configuration of the present embodiment. The configuration of the present embodiment may be implemented by adding a torque application mechanism 98 to the configuration of the third embodiment described above. It should be noted that in FIG. 19, to common elements with the elements illustrated in FIG. 1 or FIG. 15, common reference numbers are assigned, and their explanation is omitted or simplified.

The torque application mechanism 98 includes a cam 100. A cam 100 is attached to the axial shaft 44 of the balancing shaft. The cam 100 has a cam projection 106 that rotates with the axial shaft 44 of the balancing shaft.

The torque application mechanism 98 also includes spring elements 102 and 104. The spring elements 102 and 104 are arranged so that they transmit spring compressive forces to the lateral surface of the cam 100 in positions that are 180 (degrees) out of phase with each other.

The configuration of the present embodiment includes a balancing device 78, as in the case of the third embodiment. In balancing device 78, conditions arise in which no torque is transmitted to balancing shaft 80 at the first change point and the second change point. A torque application mechanism 98 is provided for applying torque in the opposite direction with respect to the direction of rotation of the crankshaft 16 to the balancing shaft 80 in these states.

FIG. 20 is a diagram for explaining a configuration operation of the present embodiment. Additionally, FIG. 21 illustrates the absolute amount of torque from the spring element that the cam 100 receives from the spring elements 102 and 104 in situations (1) to (8) illustrated in FIG. 20. It should be noted that in FIG. 21, the region in which the torque from the spring element is negative is the region in which the spring elements 102 and 104 are compressed, and the region in which the torque from the spring element is positive is the region in which the spring elements 102 and 104 are released.

In FIG. 20 and 21, in the section from (1) to (2), the axial force 108 of the connecting rod 36 generates a moment 110 that rotates the balancing shaft 80, and an offset 112 in the compression direction occurs in the spring element 102 with the rotation of the cam 100. In the section from ( 2) to (4), an offset 112 in the release direction occurs in the spring member 102, and a moment 110 is transmitted to the balancing shaft 80 through the cam 100. The configuration of this embodiment is designed so that the first change point is included in section (2) to (four). If torque 110 is transmitted to balancing shaft 80 in this section, balancing shaft 80 can forcibly maintain reverse rotation stability.

The cam 100 of the present embodiment is formed so that stretching and compression does not occur in the spring element 102 in the section (4) to (5). Therefore, the balancing shaft 80 maintains the reverse rotation mainly by the axial force of the connecting rod 36 in this section. In the section (5) to (8), the moment 110, which is almost the same as the moment 110 in the section (1) to (4), arises due to the function of the spring 104. The configuration of this embodiment is designed so that the second change point is turned on to section (6) to (8), in which the compression force of the spring generates a moment 110. Therefore, according to the configuration of the present embodiment, the balancing shaft 80 can stably rotate in the opposite direction relative to the crankshaft 16 in the entire rotation region.

Modification of the Fourth Embodiment

In this regard, in the above fourth embodiment, the section (4) to (5) and the section (8) to (1) illustrated in FIG. 20 and 21 are defined as sections in which no torque is generated from the spring element. However, the configuration of the present invention is not limited to this. In these sections, the axial force of the connecting rod 36 generates a large moment, and therefore, to smooth the actuating torque, these sections can be used as sections for compressing the spring elements 102 and 104, as illustrated in FIG. 22.

Further, in the above fourth embodiment, a desired torque is generated both in the vicinity of the first change point and in the vicinity of the second change point by using two spring elements 102 and 104. However, the configuration of the present invention is not limited thereto. FIG. 23 illustrates a configuration using a cam 114 having two cam cam projections. According to such a configuration, both at the first point of change and at the second point of change, the spring element 102 alone can forcibly generate the required torque.

Additionally, in the above fourth embodiment, the cam 100 is attached to the axial shaft 44 of the balancing shaft to form a torque. However, the configuration of the present invention is not limited to this. FIG. 24 illustrates an example in which two cam mechanisms 116 and 118 are provided on the side of a stationary member, such as a cylinder block. The cam mechanisms 116 and 118, respectively, have the function of converting the rotation operation of the connection point 38 to the balancing shaft into torque from the spring element. According to such a configuration, a torque can be appropriately generated for forcing, by means of the balancing shaft 80, stable reverse rotation, as in the case of the fourth embodiment.

Additionally, in the above fourth embodiment, in order to apply the required torque to the balancing shaft 80, the cam 100 and the spring elements 102 and 104 are again included in the internal combustion engine 10. However, the configuration of the present invention is not limited to this. For example, a direct-injection inline cylinder gasoline engine and a diesel engine include high-pressure fuel injection pumps. A high-pressure fuel injection pump may include a cam that operates in synchronization with the duty cycle of the internal combustion engine, and a spring element that engages with the cam. A torque generated in the fourth embodiment may be generated by using an existing cam, a spring element, and the like.

Further, in the above fourth embodiment, the torque application mechanism 98 is included in the configuration of the third embodiment. The balancing device 78 of the third embodiment makes it easier to self-lock at the first change point and the second change point, compared with the balancing device 30 in the first embodiment and the balancing device 66 in the second embodiment. Therefore, the configuration of the fourth embodiment shows particularly high applicability in the case of basing on the configuration of the third embodiment. However, the combined condition is not limited to this. Thus, the torque application mechanism 98 in the fourth embodiment can be combined with the configuration of the first embodiment or the second embodiment.

Fifth Embodiment

A fifth embodiment of the present invention will now be described with reference to FIG. 25-27. FIG. 25 is a diagram for explaining a configuration of the present embodiment. The configuration of the present embodiment may be implemented by adding a torque application mechanism 120 to the configuration of the third embodiment described above. It should be noted that, in FIG. 25, common elements with the elements illustrated in FIG. 1 or FIG. 15, common reference numbers are assigned, and their explanation is omitted or simplified.

In the present embodiment, the torque application mechanism 120 includes a first spring element 122. The first spring element 122 includes a spring element 124 and a contact part 126. The contact part 126 is provided so that it comes into contact with the balancing connection point 38 the shaft at the first point of change at which the axial line of the connecting rod 36 is superimposed on the center of the axial shaft 44 of the balancing shaft. Additionally, the spring member 124 is provided so as to generate a compression force of the spring that pushes the connection point 38 to the balancing shaft counterclockwise in FIG. 24 if possible.

The torque application mechanism 120 also includes a second spring element 128. The second spring element 128 is provided at a position substantially 180 (degrees) out of phase with the first spring element 122. The second spring element 128 includes a spring element 132 and a contact portion 130. The contact portion 130 is provided in such a way that it contacts the balancing shaft connection point 38 at a second change point indicating a state in which the balancing shaft 80 rotates 180 (degrees) from the first measurement point eniya. Additionally, the spring member 132 is provided to generate a compression force of the spring that pushes the connection point 38 to the balancing shaft in a counterclockwise direction in FIG. 25 if possible.

FIG. 26 is a diagram for explaining a configuration operation of the present embodiment. Additionally, FIG. 27 illustrates the absolute value of the pushing force in the x-axis direction received by the connection point 38 with the balancing shaft in situations (1) to (8) illustrated in FIG. 26. As illustrated in FIG. 26, in section (1) to (2), an axial force 134 that provides an offset 136 of the first spring element 122 in the compression direction occurs when the crankshaft 16 rotates. In section (3) to (4), an offset of 136 the direction of passage occurs in the first spring element 122. Then, the connection point 38 with the balancing shaft continues to receive the reaction force 138 in the negative direction along the x axis over almost the entire section from (1) to (4).

When the connection point 38 with the balancing shaft is located in the upper side relative to the center of the axial shaft 44 of the balancing shaft, the reaction force 138 in the negative direction along the x axis generates a moment that rotates the balancing shaft 80 in the opposite direction relative to the crankshaft. In the present embodiment, the balancing device 78 is designed so that the first change point is included in the section (1) to (4). Therefore, according to the device 78, the balancing shaft 80 can stably rotate in the reverse rotation direction in the vicinity of the first change point.

In FIG. 26, in section (5) to (8), the second spring element 128 pushes the connection point 38 with the balancing shaft in the positive x-axis direction. When the connection point 38 with the balancing shaft is located in the lower side relative to the center of the axial shaft 44 of the balancing shaft, the force in the positive direction along the x axis, which acts on the connection point 38 with the balancing shaft, generates a moment that rotates the balancing shaft 80 in the opposite direction relative to the direction rotation of the crankshaft 16. In the present embodiment, the balancing device 78 is designed so that the second change point is included in the section from (5) to (8). Therefore, according to the device 78, the balancing shaft 80 can stably rotate in the reverse rotation direction in the vicinity of the second change point.

As described above, according to the configuration of the present embodiment, the balancing shaft 80 can also stably rotate in the opposite direction relative to the crankshaft 16, as in the case of the fourth embodiment. Therefore, according to the configuration of the present embodiment, a compact internal combustion engine 10 with excellent smooth operation can also be provided.

Sixth Embodiment

Configuration of a Sixth Embodiment

Next, a sixth embodiment of the present invention is described with reference to FIG. 28-34. FIG. 28 is a diagram for explaining a configuration of a sixth embodiment of the present invention. The configuration of the present embodiment is similar to the configuration of the first embodiment (see FIG. 1) with the exception that the balancing device 30 is replaced with the offset balancing device 140.

In the present embodiment, the crankshaft 16 of the internal combustion engine 10 is used by the offset crank method. In FIG. 28, the dashed line shown by assigning the reference number 142 is the center line of the reciprocating motion of the piston 12. In the internal combustion engine 10, the center line 144 of the crankshaft 16 is located at a position separated from the center line 142 of the piston 12 by a distance H parallel with centerline 142.

The effect through an offset crank

FIG. 29 is a diagram for explaining an effect obtained by an offset crank method. The drawing on the left side in FIG. 29 shows the construction of a comparative example. In the design, the center of the main shaft 22 of the crankshaft is superimposed on the center line 142 of the reciprocating motion of the piston 12. The drawing on the right side in FIG. 29 schematically illustrates a construction using an offset crank method. In the design, the center of the main shaft 22 of the crankshaft is offset a distance H from the center line 142 of the reciprocating movement of the piston 12.

In FIG. 29, the arrow shown by assigning the reference number 146 expresses the combustion pressure that acts on the piston 12. The combustion pressure 146 reaches a large value when the piston 12 slightly passes the top dead center in the stroke. The two drawings illustrated in FIG. 29 show a state in which a large combustion pressure 146 acts on the piston 12.

The piston 12 is connected to the crank pin 18 through the connecting rod 14. Therefore, the reaction force 148 of the combustion pressure 146 is introduced into the piston 12 from the connecting rod 14. When the connecting rod 14 is tilted at δ (degrees) relative to the center line 142 of the piston 12, the reaction force 148 includes horizontal component 150 expressed by (reaction force 148 * sin δ). The horizontal component 150 acts as a force that presses the piston 12 against the side wall in the cylinder.

In the construction of the comparative example, in which the axial line 142 of the reciprocating movement of the piston 12 is superimposed on the center of the main shaft 22 of the crankshaft, a large combustion pressure 146 is applied to the piston 12 in a state in which the connecting rod 14 is inclined relative to the axial line 142. Therefore, in this configuration, a large horizontal component 150 is formed, and the friction of the piston 12 tends to be large.

According to the offset crank method, the angle of inclination δ of the connecting rod 14 during the action of the high combustion pressure 146 on the piston 12 can be set to a small value. Therefore, according to the configuration, the piston 12 can reciprocate in the cylinder without receiving the large horizontal component 150. Therefore, according to the configuration of the present embodiment, the friction of the piston 12 can be reduced and fuel efficiency is improved, compared with the configuration of the comparative example.

Asymmetry due to bias

FIG. 30 is a diagram for explaining the asymmetry of the rotational speed profile shown by the piston 12 in the present embodiment. In FIG. 30, the piston 12 is located at top dead center at an angle of rotation of the crankshaft θ = 0 (° CA) and is located at bottom dead center at θ = 180 (° CA).

Regarding the angle of rotation θ of the crankshaft, θ = 90 (° CA) and θ = 270 (° CA) are the midpoints between top dead center and bottom dead center. However, between the angle of inclination δ90 of the connecting rod 14 at θ = 90 (° CA) and the angle of inclination δ 270 at θ = 270 (° CA), significant differences arise due to the influence of the offset crank. As a result, the piston stroke PS90 that occurs when the crank angle θ = 90 (° CA) and the piston stroke PS270 that occurs when θ = 270 (° CA) have different meanings. Necessarily, the course (PS180-PS90) that occurs when θ changes from 90 to 180 (° CA) and the course (PS270-PS180) that occurs when θ changes from 180 to 270 (° CA) also have different meanings.

As described above, in the internal combustion engine 10, which uses the offset crank method, the piston 12 exhibits asymmetric displacement profiles in the external process with the passage of the angle θ of rotation of the crankshaft from top dead center to bottom dead center and in the opposite way with passage from bottom dead center to top dead center. When displacement profiles are asymmetric, the inertia forces that occur with displacements also become asymmetric. Therefore, in order to balance the inertia generated by the piston 12 of the internal combustion engine 10 with high accuracy, it is effective to cause the balancing device 140 to generate an asymmetric vibration-inducing force.

Balancing Device Configuration of the Present Embodiment

FIG. 31A illustrates the configuration of the balancing device 140 that is used in the present embodiment. The balancing device 140 is similar to the balancing device 30 (see FIG. 2) in the first embodiment, with the exception that the center line 56 is “crankshaft - balancing shaft”, which connects the center of the main shaft 22 of the crankshaft and the axial shaft 44 of the balancing shaft , is offset so that it is separated from the center line 152 of the guide section 54 by a distance h. Further in this document, in FIG. 31A, common elements with elements illustrated in FIG. 2, common references with numbers are assigned, and their explanation is omitted or simplified.

In the balancing device 140 illustrated in FIG. 31A, the sliding portion 52 of the connecting rod 36 performs a linear motion in the guide section 54 with rotation of the crankshaft 16. When moving, the hinge 50 of the connecting rod 36 moves along the center line 152 from the guide section 54. If the hinge 50 moves on the center line 56 "crankshaft - balancing shaft ", the angle of rotation α of the balancing shaft must show symmetrical profiles of changes in the process in an external way, in which the angle θ of rotation of the crankshaft changes from zero to 180 (° CA), and in the opposite way, in which θ changes It is a C of 180 to 360 (° CA). However, when the hinge 50 moves on the axial line 152, which deviates from the center line 56 "crankshaft - balancing shaft", the profiles of the angle α of rotation of the balancing shaft become asymmetric in the process of the external method and in the opposite way.

The asymmetry of the angle α of rotation of the balancing shaft, as described above, occurs because the hinge 50 moves on the axial line 152, deviated from the Central line 56 "crankshaft - balancing shaft". FIG. 31B, 31C and 31D illustrate examples of other balancing devices in which such asymmetry occurs. More specifically, FIG. 31B illustrates a balancing device 154 in which the center of the axial shaft 44 of the balancing shaft is offset a distance h from the center line 152 from the guide section 54. FIG. 31C illustrates a balancing device 156 in which the center of the main shaft 22 of the crankshaft is offset a distance h from the center line 152 from the guide section 54. Additionally, FIG. 31D illustrates a balancing device 158 in which both the center of the main shaft 22 of the crankshaft and the center of the axial shaft 44 of the balancing shaft are offset by a distance h, respectively, in opposite directions from the center line 152. These balancing devices 154, 156 and 158 can be used by appropriate replacement instead of the balancing device 140 illustrated in FIG. 31A, in accordance with the asymmetry of the inertia force that is generated by the piston 12.

Vibration causing force generated by balancing device of an embodiment of the invention

Hereinafter, with reference to FIG. 32-34, as well as FIG. 31A, a vibration causing force which is generated by the balancing device 140 of the present embodiment is described. The different sizes of the balancing device 140 become different parameters that have an effect on the profile of the angle of rotation α of the balancing shaft. Here, the simulation result is described. The simulation is performed by setting the ratio r1 / 1cb of the radius r1 of rotation of the point 32 of the connection with the crankshaft and the length 1cb of the connecting rod 36 is fixed equal to the normal value and changing the ratio h / r1 of the displacement value h and the radius of rotation r1 to "-a", "-b" and "+ a". It should be noted that the signs “+” and “-”, assigned by “a” and “b,” express differences in the direction in which the offset by the distance h is performed.

FIG. 32 illustrates the relationship between the crank angle θ and the angular velocity dα / dθ of the balancing shaft. Since the balancing shaft 40 rotates in the opposite direction relative to the crankshaft 16, the angular velocity dα / dθ of the balancing shaft becomes -1 if α and θ exhibit an identical change. For example, in the form of the “-a” signal, such an asymmetry appears that the change in α is small outside θ relative to θ, passing from top dead center to bottom dead center, and a trend opposite to the above appears in the opposite way from bottom dead center to top dead point. Additionally, in the “-b” waveform, less asymmetry than the “-a” waveform appears. Additionally, in the form of the “+ a” signal, an asymmetry opposite to the “-a” asymmetry appears.

As described above, the balancing device 140 of the present embodiment can generate an angular velocity dα / dθ of the balancing shaft, which has asymmetric profiles in an external and reverse manner. A centrifugal force proportional to the square of the angular velocity and the reaction force to the angular velocity acts on the balancing shaft 40. The balancing shaft 40 generates a vibration causing force corresponding to the composite value of the centrifugal force and the reaction force. In the present embodiment, the angular velocity dα / dθ of the balancing shaft has asymmetric profiles, and therefore the vibration-induced force generated by the balancing shaft 40 also has asymmetric profiles in an external and reverse manner, similar to the inertia force of the piston 12.

FIG. 33 is a diagram showing a vibration causing force obtained by combining the Y axis component of the vibration causing force generated by the crankshaft 16 and the Y axis component of the vibration causing force generated by the balancing shaft 40, compared to the inertia of the piston 12. Five types of waveforms illustrated in FIG. 33, respectively, are described as follows:

Signal form 60 * 2: corresponds to a two-fold excess of the component along the Y axis of the vibration-causing force generated by the crankshaft 16 (see Fig. 8)

Waveform 160: asymmetric inertia generated by piston 12

Waveform 162: corresponds to a component object of the component along the Y axis of the vibration causing force generated by the balancing shaft 40 under the condition “-a”, and the component along the Y axis of the vibration causing force generated by the crankshaft 16

Waveform 164: corresponds to a component object of the y-axis causing the vibration force generated by the balancing shaft 40 under the condition “-b”, and the y-axis component of the vibration causing force generated by the crankshaft 16

Waveform 166: corresponds to a component object of the component along the Y axis of the vibration causing force generated by the balancing shaft 40 under the condition “+ a”, and the component along the Y axis of the vibration causing force generated by the crankshaft 16

The vibration-causing force generated by the crankshaft 16 is in the form of a sine wave with little or no distortion with respect to the change in the angle of rotation θ of the crankshaft. If the balancing shaft 40 rotates at an equal rotational speed relative to the rotational speed of the crankshaft 16, the vibrational forces generated by both of them should be almost identical. In this case, the inertia force generated by the balancing device 140 should show almost symmetrical profiles in an external way and in the opposite way, as shown by the waveform 60 * 2. Waveforms 162, 164, and 166 have asymmetric profiles. Since the inertia generated by the piston 12 has an asymmetry, it is believed that these waveforms 162, 164 and 166 have a higher ability than the waveform 60 * 2 with respect to balancing the inertia of the piston 12.

FIG. 34 illustrates waveforms of unbalanced forces remaining in an internal combustion engine 10 as a result of combining various types of vibration causing forces illustrated in FIG. 33, with the inertia of the piston 12. The semantic meanings of the waveforms illustrated in FIG. 34, respectively, are described as follows:

Waveform 160 + 162: unbalanced force remaining in the internal combustion engine 10 under condition “-a”

Waveform 160 + 164: unbalanced force remaining in the internal combustion engine 10 under condition “-b”

Waveform 160 + 166: unbalanced force remaining in the internal combustion engine 10 under the condition "+ a"

Waveform 160 + 60 * 2: unbalanced force remaining in the internal combustion engine 10 when the balancing shaft 40 is controlled by a gear train

As illustrated in FIG. 34, under any of the conditions according to the balancing device 140 of the present embodiment, the remaining unbalanced force may be reduced compared to the case of controlling the balancing shaft 40 by means of a gear transmission. In particular, when simulating this time, in the case of using the “-b” condition, the remaining unbalanced force can be set equal to a sufficiently small value (see waveform 160 + 164). Thus, according to the configuration of the present embodiment, even if the offset crank method is used, excellent smoothness of operation can also be provided for the internal combustion engine 10 using a compact mechanism using a connecting rod 36.

Seventh Embodiment

Seventh Embodiment Configuration

Next, a seventh embodiment of the present invention is explained with reference to FIG. 35-38.

FIG. 35 is a diagram for explaining a configuration of a balancing device 168 that is used in a seventh embodiment of the present invention. The configuration of the present embodiment is similar to the configuration of the first embodiment (see FIG. 1) or the configuration of the sixth embodiment (see FIG. 28), except for such an aspect that the balancing device 30 is replaced by the balancing device 168.

Additionally, the balancing device 168 of the present embodiment is similar to the balancing device 66 in the second embodiment (see FIG. 14), except that the center of the rotational shaft 72 of the guide section 68 is offset so that it is separated from the center line 56 " crankshaft - balancing shaft "at a distance h. Further in this document, in FIG. 35, identical or common elements with the elements illustrated in FIG. 14, identical reference numbers are assigned, and their explanation is omitted or simplified.

The balancing device 168 of the present embodiment is adapted to the offset crank method, similarly to the balancing device 140 of the sixth embodiment. In other words, the balancing device 168 illustrated in FIG. 35 rotates the balancing shaft 40 asymmetrically at an unequal speed in a process in which the angle θ of rotation of the crankshaft changes from zero to 180 (° CA), and in a process in which θ changes from 180 to 360 (° CA), similar to a balancing device 140 in a sixth embodiment.

If the center of the rotary shaft 72 of the guide section 68 is superimposed on the crankshaft-balancing shaft center line 56 in the balancing device 168 illustrated in FIG. 35, the angle of rotation α of the balancing shaft should show symmetrical change profiles on the positive side (side 90 (° CA)) and on the negative side (side 270 (° CA)) with the angle of rotation of the crankshaft θ = 0 (° CA) as the boundary . However, when the center of the rotational shaft 72 deviates from the center line 56 "crankshaft - balancing shaft", the above symmetry is broken, and the profiles of the angle of rotation α of the balancing shaft become asymmetric.

FIG. 36, 37 and 38 suitably correspond to FIG. 32, 33 and 34 in the above sixth embodiment. The explanations of these drawings are common to the explanations of the respective drawings, and therefore, an additional explanation is omitted here.

As illustrated in FIG. 36-38, by means of the balancing device 168 in the present embodiment, an asymmetric vibration-inducing force which is suitable for balancing the asymmetric inertia force can be generated, as in the case of the sixth embodiment. Therefore, according to the configuration of the present embodiment, excellent smoothness of operation can also be provided for the internal combustion engine 10 using the offset crank method, without the difficulty of reducing size and weight, as in the case of the sixth embodiment.

Eighth Embodiment

Configuration of Eighth Embodiment

Next, an eighth embodiment of the present invention is explained with reference to FIG. 39-45.

FIG. 39 illustrates the configuration of an internal combustion engine on which a balancing device 170 of an eighth embodiment of the present invention is mounted. The balancing device 170 has a connecting rod 36, which is located by tilting relative to the axial line 172 of the reciprocating movement of the piston 12, as in the cases of the first to seventh embodiments described above. Here, the angle of inclination of the centerline 172 of the piston 12 and the connecting rod 36 is presumably “β” degrees. Additionally, a straight line that is parallel with the axial line 172 of the piston 12 and passes through the main shaft 22 of the crankshaft, and a straight line that passes through the axial shaft 44 of the balancing shaft, respectively, are referred to as "the axial line 174 of the crankshaft" and " center line 176 of the balancing shaft. "

The balancing device 170 of the present embodiment includes a crankshaft 178 and a balancing shaft 180. The crankshaft 178 and the balancing shaft 180, respectively, have an eccentric load 182 of the crankshaft and an eccentric load 184 of the balancing shaft. The balancing device 170 of the present embodiment is similar to the balancing device 30 (see FIG. 1) of the first embodiment except that the eccentric crankshaft load 182 and the balancing shaft eccentric load 184 include centers of gravity 186 and 188, respectively, in positions shown by dots in the form of a black circle in FIG. 39.

FIG. 40 is a diagram for explaining the principle of how the balancing device 170 of the present embodiment balances the vibrational forces that arise in an internal combustion engine. In particular, the left side in FIG. 40 is a drawing illustrating conditions that an eccentric crankshaft load 182 and a balancer shaft eccentric load 184 must meet to balance the vibrational force Fp caused by the piston 12 and the vibrational force Fc caused by the connecting rod 14. Further, the right side on FIG. 40 is a drawing illustrating the conditions that an eccentric crankshaft load 182 and a balancer shaft eccentric load 184 must meet in order to balance the vibrational force Fc caused by the connecting rod 36. Hereinafter, similar to the explanation in the first embodiment, the weight of the piston 12 expressed by "mp", and the weight of the connecting rod is expressed by "mc". Additionally, in the present embodiment, the weight of the connecting rod 36 is expressed by “mr”.

The two drawings illustrated in FIG. 40, express the state of time at which the piston 12 reaches top dead center. In this case, the piston 12 and the connecting rod 14 generate a vibration causing force (FCS + FBS) corresponding to “mp + mc” in the upward direction in FIG. 40 along the center line 172 of the reciprocating motion of the piston 12. In order to balance the vibration causing force (FCS + FBS), the eccentric crankshaft load 182 and the balancing shaft eccentric load 184 must generate a vibration causing force with an identical absolute value in the downward direction in FIG. 40.

In the present embodiment, in order to respond to the above requirement, the eccentric crankshaft load 182 is forcibly responsible for the downward force causing vibration in FIG. 40, which corresponds to “mc + mp / 2”, as illustrated on the left side in FIG. 40. Additionally, the eccentric load of the balancing shaft is forcibly responsible for the downward force causing vibration in FIG. 40, which corresponds to "mp / 2". According to this setting, the vibration causing force (FCS + FBS), which is generated by the piston 12 and the connecting rod 14 at top dead center, can be properly balanced.

In a situation in which the piston 12 reaches bottom dead center, the piston 12 and the connecting rod 14 generate a vibration-causing force (FCS + FBS) with a direction inverted with respect to the top dead center case. At this time, the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft also generate vibrational forces with inverted directions. Therefore, according to the above setting, vibration causing forces can also be balanced by each other at bottom dead center.

Between the top dead center and the bottom dead center, the piston 12 moves only vertically along the axis of reciprocating movement of the piston 12 and does not generate a vibration force other than the vibration force in the direction of the axial line 172. The connecting rod 14 rotates around the main shaft 22 of the crankshaft shaft and generates a vibrational force Fc corresponding to the centrifugal force of weight mc. At this stage, the eccentric load 182 of the crankshaft, which rotates similarly, generates a vibration-causing force corresponding to the centrifugal force of the weight "mc + mp / 2". The vibration causing force Fc caused by the connecting rod 14 is balanced by the part corresponding to the “mc” vibration causing force of the crankshaft eccentric load 182. Additionally, the remainder of the vibration causing force generated by the crankshaft eccentric load 182, i.e. the part corresponding to "mp / 2" is balanced by the vibration-causing force of the eccentric load 184 of the balancing shaft, which rotates in the opposite direction relative to the eccentric load 182 of the crankshaft. Thus, according to the condition illustrated on the left side in FIG. 40, the vibration-causing force (FCS + FBS) accompanying the movement of the piston 12 and the connecting rod 14 can always be balanced by the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft.

The drawing on the right side of FIG. 40 illustrates a state in which the connecting rod 36 reaches the working end in the upper right side of FIG. 40 in synchronization with piston 12 reaching top dead center. In this state, the connecting rod 36 generates a vibration causing force Fr to the upper right corner in FIG. 40 with an inclination angle of β degrees relative to the center line of the crankshaft 174. In the present embodiment, in order to balance the vibration causing force Fr, the eccentric crankshaft load 182 and the balancing shaft eccentric load 184 forcefully generate vibrational forces directed to the lower left side of FIG. 40, each of which corresponds to a weight of "mr / 2".

In the situation illustrated in FIG. 40, the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft, respectively, generate vibrational forces corresponding to a weight of "mr / 2". The vibration-causing forces extend to the lower left side of FIG. 40 at an angle of β degrees relative to the axial line 174 of the crankshaft and the axial line 176 of the balancing shaft. According to the vibration causing forces, the vibration causing force Fr caused by the connecting rod 36 can be appropriately balanced.

When the piston 12 reaches bottom dead center, the connecting rod 36 generates a vibrational force Fr corresponding to the weight mr to the lower left side in FIG. 40. At this time, the eccentric load of the crankshaft 182 and the eccentric load of the balancing shaft 184 generate vibrational forces in the upward direction to the right in FIG. 40. Therefore, according to the above setting, the vibration causing force caused by the connecting rod 36 can also be balanced at the bottom dead center.

Between the top dead center and the bottom dead center, the connecting rod 36 does not generate a large vibration causing force. On the other hand, the weight "mr / 2" that is provided to the eccentric load 182 of the crankshaft and the weight "mr / 2" that is provided to the eccentric load 184 of the balancing shaft generate vibrational forces when rotated in opposite directions, respectively. These forces causing vibration are in opposite directions relative to each other and, therefore, are balanced.

When the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft satisfy the conditions shown on the right side in FIG. 40, the vibration-causing force accompanying the movement of the connecting rod 36 can always be balanced by the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft.

In the present embodiment, the conditions illustrated on the left side in FIG. 40, and the conditions illustrated on the right side are combined and applied to the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft, respectively. In particular, the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft, respectively, are provided with such weights and centers of gravity that satisfy the conditions described as follows.

Crankshaft eccentric load

The eccentric load of the crankshaft generates a vibration causing force, which is obtained by combining two vibration causing forces having absolute values and directions described as follows when the piston 12 reaches top dead center.

(1) Absolute value: corresponds to weight (mc + mp / 2)

Direction: The opposite direction with respect to the vibrating force Fp (hereinafter referred to as the “direction opposite to Fp”)

(2) Absolute value: corresponds to weight (mr / 2)

Direction: The opposite direction with respect to the vibrating force Fr (hereinafter referred to as the “direction opposite to Fr”)

Eccentric load balancing shaft

The eccentric load of the balancing shaft generates a vibration causing force which is obtained by combining two vibration causing forces having absolute values and directions described as follows when the piston 12 reaches top dead center.

(1) Absolute value: corresponds to weight (mp / 2)

Direction: Direction opposite to Fp

(2) Absolute value: corresponds to weight (mr / 2)

Direction: Direction opposite to Fr

FIG. 41 is an enlarged view of a crankshaft 178 illustrated in FIG. 39, i.e. crankshaft 178, which is used in the present embodiment. The crankshaft 178 has a weight and a center of gravity that satisfy the conditions described above. In particular, the crankshaft 178 has a weight of almost (mc + mp / 2 + mr / 2) and has a center of gravity 186 at the position illustrated by the black circle dot in FIG. 41.

If the crankshaft 178 is set so that it balances only the vibrational force (Fp + Fc) caused by the piston 12 and the connecting rod 14, the center of gravity position 186 at top dead center should preferably overlap the crankshaft center line 174. In contrast, in the present embodiment, in order to balance the vibration causing force Fr caused by the connecting rod 36, the position of the center of gravity 186 is shifted to the direction opposite to Fr by βcs degrees from the crankshaft center line 174. Thus, the crankshaft 178 of the present embodiment has a center of gravity 186 on the opposite side of the connection point 32 with the crankshaft relative to the crankshaft center line 174. It should be noted that the shear angle in βcs degrees of the center of gravity 186 necessarily becomes an angle smaller than the angle of inclination in β degrees of the connecting rod 36.

FIG. 42 is a perspective view of the crankshaft 178. Here, the connecting rod 36 is supposedly connected on the left side of the crankshaft 178. The crankshaft 178 has two eccentric weights 182 of the crankshaft with a crank pin 18 between them. The vibration-causing force (FCS + FBS) of the piston 12 and the connecting rod 14 acts on the crank pin 18. Therefore, the weight “mc + mp / 2”, which balances the force causing vibration, is preferably evenly distributed in two crankshaft eccentric weights 182. On the other hand, the vibrational force Fr caused by the connecting rod 36 acts on the vicinity of the left end in FIG. 42 from the crankshaft 178. If the weight to balance the vibrational force Fr is applied to the eccentric load 182 of the crankshaft on the right side in FIG. 42, the vibration causing force Fr, which is generated by the connecting rod 36, and the vibration-causing force, which is generated by the crankshaft eccentric load 182 on the right side in FIG. 42 should provide great torque to the crankshaft 178. Therefore, the weight “mr / 2” to balance the vibrational force Fr caused by the connecting rod 36 is preferably provided to the eccentric load 182 of the crankshaft on the side near the connecting rod 36.

In the present embodiment, in response to the above requirement, only half the weight “mc + mp / 2” is provided to the eccentric crankshaft load 182 on the right side of FIG. 42. It should be noted that the center of gravity directly of this eccentric crankshaft load 182 is present at a position that overlays the crankshaft center line 174 at top dead center. The eccentric load 182 of the crankshaft on the left side in FIG. 42 provides all the weight "mr / 2" in addition to half the weight "mc + mp / 2". It should be noted that the weight “mr / 2” is provided to the eccentric load 182 of the crankshaft on the left side in FIG. 42, so that the centers of gravity of the two eccentric weights 182 of the crankshaft become the center of gravity 186 illustrated in FIG. 41. According to this setting, the vibrational force Fr caused by the connecting rod 36 can be balanced without causing a large moment that acts on the crankshaft 178.

FIG. 43 is an enlarged view of the balancing shaft 180 illustrated in FIG. 39, i.e. balancing shaft 180 used in the present embodiment. Here, the connecting rod 36 is supposedly located on the left side of the balancing shaft 180. The eccentric load 184 of the balancing shaft of the balancing shaft 180 has a large diameter part 190 and a small diameter part 192. Part 190 with a large diameter is provided in the vicinity of one end 194 on the side that should be connected to the connecting rod 36. Part 192 with a small diameter is provided on the side of the other end 196 of the balancing shaft 180.

The eccentric load 184 of the balancing shaft is provided with a weight of almost "mp / 2 + mr / 2". From the weight, the weight "mp / 2" to balance the vibrational force Fp caused by the piston 12 is distributed equally to the large diameter part 190 and the small diameter part 192. On the other hand, a weight “mr / 2” for balancing the vibrational force Fr caused by the connecting rod 36 is provided only to the larger diameter part 190. As a result, the larger diameter part 190 has a larger outer diameter than the small diameter part 192.

The center of gravity 188 of the eccentric load 184 of the balancing shaft is provided on the opposite side of the connection point 38 with the balancing shaft with the axial line 176 of the balancing shaft between the connection point 38 with the balancing shaft and the center of gravity 188, as in the case of the crankshaft 178 (see Fig. 39) . In particular, the center of gravity 188 of the eccentric load 184 of the balancing shaft deviates from the axial line 176 of the balancing shaft by a fixed angle in the direction of inclination of the connecting rod 36. The angle of deviation necessarily becomes a smaller angle than the angle of β degrees of the connecting rod 36. According to this setting, the force causing vibration Fr caused by the connecting rod 36 can be balanced without causing a large moment, which acts on the balancing shaft 180.

FIG. 44 is a diagram for explaining an operation of an internal combustion engine loaded with a balancing device 170 of the present embodiment. In particular, FIG. 44 expresses the state of the internal combustion engine at intervals of 45 (° CA) from 0 (° CA) to 360 (° CA). In FIG. 44, the vibration causing force generated by the eccentric load 182 of the crankshaft is referred to as FCS, and the vibration causing force generated by the eccentric load 182 of the balancing shaft is referred to as FBS. It should be noted that the vibrational forces FCS and FBS are expressed by decomposing into two vibrational forces, which are the vibrational force to balance (Fp + Fc) and the vibrational force to balance Fr, to facilitate understanding. As illustrated in FIG. 44, according to the configuration of the present embodiment, not only the vibrational force (Fp + Fc) caused by the piston 12 and the connecting rod 14, but also the vibrational force Fr caused by the connecting rod 36, can always be balanced properly.

An example of a modification of an eighth embodiment of the invention

FIG. 45 is a perspective view of another example of a balancing shaft that is applicable to the balancing device 170 of the present embodiment. The balancing shaft 198 illustrated in FIG. 45 has an eccentric load 200 of a balancing shaft. The eccentric load 200 of the balancing shaft is formed so that the outer diameter gradually becomes smaller from one end 194 on the side to be connected to the connecting rod 36, to the other end 196. According to this configuration, the weight mr to balance the vibrational force Fr caused by the connecting rod 36 may be reflected on the side of one end 194 more significantly than the side of the other end 196, as is the case with the balancing shaft 180 illustrated in FIG. 43. Consequently, according to the balancing shaft 198 illustrated in FIG. 45, the vibrational force Fr can also be balanced without generating a large moment.

In this regard, in the above eighth embodiment, the weight for balancing the vibrational force Fr caused by the connecting rod 36 is reflected significantly on the side near the connecting rod 36. However, the feature is not essential for the present invention. Thus, the weight (mr / 2) that is provided to the crankshaft 178 to balance the vibrational force Fr can be evenly distributed into the two eccentric weights 182 of the crankshaft. Similarly, for the balancing shaft 180, the weight (mr / 2) can be evenly distributed over the entire area of the balancing shaft 180.

Further, in the above eighth embodiment, the configuration for balancing the vibration causing force Fr caused by the connecting rod 36 is included in the balancing device in the first embodiment, but the present invention is not limited thereto. Thus, a configuration for balancing the vibrational force Fr in any of the second to seventh embodiments can be included.

Further, in the above eighth embodiment, the weight for balancing the vibrational force Fr caused by the connecting rod 36 is uniformly reflected in the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft, but the present invention is not limited thereto. Thus, the weight for balancing the vibrational force Fr can be unevenly reflected in the eccentric load 182 of the crankshaft and the eccentric load 184 of the balancing shaft. This aspect likewise applies to the ninth embodiment, which is described below.

Ninth Embodiment

Next, a ninth embodiment of the present invention is explained with reference to FIG. 46-51.

FIG. 46 illustrates the configuration of an internal combustion engine on which a balancing device 202 of a ninth embodiment of the present invention is mounted. The balancing device 202 includes a crankshaft 204 and a balancing shaft 206. The crankshaft 204 and the balancing shaft 206 have an eccentric load 208 of the crankshaft and an eccentric load 210 of the balancing shaft, respectively. The balancing device 202 of the present embodiment is similar to the balancing device 170 (see FIG. 39) of the eighth embodiment except for the following two aspects.

(1) A connection point 32 with the crankshaft is provided on the weight side of the eccentric load 208 of the crankshaft relative to the main shaft 22 of the crankshaft, and a connection point 38 with the balancing shaft is provided on the weight side of the eccentric load 210 of the balancing shaft relative to the axial shaft 44 of the balancing shaft.

(2) The eccentric weight of the crankshaft 208 and the eccentric weight of the balancing shaft 210 include centers of gravity 212 and 214 at the positions shown by the black circle dots in FIG. 46, respectively.

FIG. 47 is a diagram for explaining the principle of the balancing device 202 of the present embodiment, balancing the vibration causing forces that are generated in the internal combustion engine. The left side in FIG. 47 is a drawing illustrating the conditions that the crankshaft eccentric weight 208 and the balancing shaft eccentric weight 210 must satisfy in order to balance the vibrational force (Fp + Fc) caused by the piston 12 and the connecting rod 14. The conditions are almost similar to the conditions explained in the eighth with reference to the left column in FIG. 40.

The two drawings illustrated in FIG. 47, each express a state at the time of reaching the top dead center by means of the piston 12. In the configuration of the present embodiment, the connecting rod 36 slides down and to the left from the top to the right of FIG. 47 in the process of moving the piston 12 from bottom dead center to top dead center. In the situation of top dead center, the connecting rod 36 reaches the movable end in the lower left side in FIG. 47.

The drawing on the right side in FIG. 47 illustrates a state in which the connecting rod 36 reaches the movable end in the lower left side of FIG. 47 in synchronization with piston 12 reaching top dead center. In this state, the connecting rod 36 generates a vibration causing force Fr to the lower left side in FIG. 47 with an inclination angle of β degrees relative to the center line of the crankshaft 174. In the present embodiment, in order to balance the vibration causing force Fr, the eccentric crankshaft load 208 and the balancing shaft eccentric load 210 force the vibration causing forces to the upper right side in FIG. 47, which correspond to a weight of "mr / 2", respectively, in a top dead center situation.

In the situation illustrated in FIG. 47, the eccentric load 208 of the crankshaft and the eccentric load 210 of the balancing shaft, respectively, generate vibrational forces corresponding to a weight of "mr / 2". These vibrational forces are directed upward to the right in FIG. 47 at an angle of β degrees relative to the axial line 174 of the crankshaft and the axial line 176 of the balancing shaft. According to these vibration causing forces, the vibration causing force Fr caused by the connecting rod 36 can be appropriately balanced.

When the piston 12 reaches bottom dead center, the connecting rod 36 reaches the movable end in the upper right side of FIG. 47 and generates a vibrational force Fr corresponding to the weight mr to the upper right side. In this case, the eccentric load 208 of the crankshaft and the eccentric load 210 of the balancing shaft generate vibrational forces to the lower left side in FIG. 47. Therefore, according to the above setting, the vibration causing force caused by the connecting rod 36 can also be balanced at the bottom dead center.

Between the top dead center and the bottom dead center, the connecting rod 36 does not generate a large vibration causing force. On the other hand, the weight "mr / 2" that is provided to the eccentric load 208 of the crankshaft and the weight "mr / 2" that is provided to the eccentric load 210 of the balancing shaft mutually balance the vibration-causing forces with corresponding rotation in opposite directions.

As described above, in the case of an eccentric load 208 of the crankshaft and an eccentric load 210 of the balancing shaft satisfying the conditions illustrated on the right side in FIG. 47, the vibration-causing force accompanying the movement of the connecting rod 36 can always be balanced by the eccentric load 208 of the crankshaft and the eccentric load 210 of the balancing shaft.

In the present embodiment, the conditions illustrated on the left side in FIG. 47, and the conditions illustrated on the right side are jointly applied to the eccentric load 208 of the crankshaft and the eccentric load 210 of the balancing shaft, respectively. In particular, weights and centers of gravity that satisfy the conditions described as follows, respectively, are provided to the eccentric load 208 of the crankshaft and the eccentric load 210 of the balancing shaft.

Crankshaft eccentric load

The eccentric load of the crankshaft generates a vibration causing force obtained by combining two vibration causing forces having absolute values and directions described as follows when the piston 12 reaches top dead center.

(1) Absolute value: corresponds to weight (mc + mp / 2)

Direction: Direction opposite to Fp

(2) Absolute value: corresponds to weight (mr / 2)

Direction: Direction opposite to Fr

Eccentric load balancing shaft

The eccentric load of the balancing shaft generates a vibration-causing force obtained by combining two vibration-causing forces having absolute values and directions described as follows when the piston 12 reaches top dead center.

(1) Absolute value: corresponds to weight (mp / 2)

Direction: Direction opposite to Fp

(2) Absolute value: corresponds to weight (mr / 2)

Direction: Direction opposite to Fr

FIG. 48 is an enlarged diagram of a crankshaft 204 illustrated in FIG. 46, i.e. crankshaft 204 for use in the present embodiment. The crankshaft 204 has a weight and a center of gravity that satisfy the conditions described above. In particular, the crankshaft 204 has a weight of almost (mc + mp / 2-mr / 2) and has a center of gravity 212 at the position shown by the black circle dot in FIG. 48.

If the crankshaft 204 only balances the vibrational force (Fp + Fc) caused by the piston 12 and the connecting rod 14, the center of gravity 212 at top dead center should preferably overlap the crankshaft center line 174. In this regard, in the present embodiment, in order to balance the vibration causing force Fr caused by the connecting rod 36, the position of the center of gravity 212 is shifted by βcs degrees to the direction direction opposite to Fr from the crankshaft center line 174. Thus, the crankshaft 204 of the present embodiment has a center of gravity 212 on the opposite side of the connection point 32 with the crankshaft relative to the crankshaft center line 174.

The crankshaft 204 includes two eccentric weights 208 of the crankshaft, similar to the crankshaft 178 illustrated in FIG. 42. In the present embodiment, the weight “Fr / 2” for balancing the vibrational force Fr caused by the connecting rod 36 is reflected only in the eccentric load 208 of the crankshaft, which is located near the connecting rod 36, as in the case of the eighth embodiment. In particular, the eccentric load 208 of the crankshaft, which is located at a position far from the connecting rod 36, is provided with half the weight “mc + mp / 2” to balance the vibrational force (Fp + Fc). On the other hand, the eccentric load 208 of the crankshaft, which is located near the connecting rod 36, is provided with a weight of "(mc + mp / 2) / 2-mr / 2", which is the result of subtracting the weight of "mr / 2" to balance the vibrational force Fr half the weight is "mc + mp / 2". A second crankshaft eccentric weight 208 is formed such that the centers of gravity of the two crankshaft eccentric weights 208 become the center of gravity 212, illustrated in FIG. 48. According to such a configuration, the vibrational force Fr caused by the connecting rod 36 can be balanced without causing a large moment that acts on the crankshaft 204, as in the case of the eighth embodiment.

FIG. 49 is an enlarged view of a balancing shaft 206 illustrated in FIG. 46, i.e. balancing shaft 206, which is used in the present embodiment. The eccentric load 210 of the balancing shaft of the balancing shaft 206 comprises a small diameter part 216 and a large diameter part 218. Part 216 with a small diameter is formed in the vicinity of one end 194 on the side that connects to the connecting rod 36. Part 218 with a large diameter is formed in the vicinity of the other end 196 of the balancing shaft 206.

The eccentric load 210 of the balancing shaft is provided with a weight of almost "mp / 2-mr / 2". From the weight, the weight of "mp / 2" is distributed equally to part 216 with a small diameter and part 218 with a large diameter. The magnitude of "mr / 2" weight reduction is reflected only in part 216 with a small diameter. The center of gravity 214 of the eccentric load of the balancing shaft 210 is provided on the opposite side of the point 38 of the connection with the balancing shaft with the axial line 176 of the balancing shaft between the point 38 of the connection with the balancing shaft, and the center of gravity 214, as in the case of the crankshaft 204 (see Fig. 46 ) According to this setting, the vibrational force Fr caused by the connecting rod 36 can be balanced without causing a large moment that acts on the balancing shaft 206.

FIG. 50 is a diagram for explaining an operation of an internal combustion engine on which a balancing device 202 of the present embodiment is mounted. In particular, FIG. 50 expresses the state of the internal combustion engine at intervals of 45 (° CA) from 0 (° CA) to 360 (° CA). In FIG. 50, the vibrational force generated by the crankshaft eccentric load 208 is expressed by FCS and (-FCS). “FCS” stands for virtual vibration-causing force corresponding to the weight of “mc + mp / 2”, and “-FCS” stands for virtual negative vibration-causing force, which is generated by the “-mr / 2” weight reduction value. In fact, the eccentric load 208 of the crankshaft generates a vibration-inducing force corresponding to the resulting vector from the FCS vector and the -FCS vector.

FIG. 50 likewise expresses a vibration-inducing force generated by the eccentric load of the balancing shaft 210 by FBS and (-FBS). “FBS” stands for virtual vibration-inducing force corresponding to the weight of “mp / 2”, “-FBS” refers to virtual negative vibration-inducing force that is generated by the “-mr / 2” weight reduction value. In fact, the eccentric load 210 of the balancing shaft generates a vibrational force corresponding to the resulting vector from the "FBS" vector and the "-FBS" vector.

As illustrated in FIG. 50, according to the configuration of the present embodiment, the vibrational force (Fp + Fc) caused by the piston 12 and the connecting rod 14 is balanced by the vibrational force (FCS + FBS) caused by the eccentric load 208 of the crankshaft and the eccentric load 210 of the balancing shaft. Further, the vibrational force Fr caused by the connecting rod 36 is balanced by the effect of (-FCS-FBS) of reducing the weight of the eccentric load 208 of the crankshaft and the eccentric load 210 of the balancing shaft. Therefore, according to the configuration of the present embodiment, the vibration-causing forces accompanying the operation of the internal combustion engine can always be properly balanced.

An example of a modification of the ninth embodiment of the invention

FIG. 51 is a perspective view of another example of a balancing shaft that is applicable to the balancing device 202 of the present embodiment. The balancing shaft 220 illustrated in FIG. 51, has an eccentric load 222 balancing shaft. The eccentric load 222 of the balancing shaft is formed so that the outer diameter gradually becomes large from one end 194 on the side that connects to the connecting rod 36 to the other end 196. According to this configuration, the weight mr to balance the vibrational force Fr caused by the connecting rod 36 , may be reflected on the side of one end 194 more significantly than the side of the other end 196, as in the case of the balancing shaft 206 illustrated in FIG. 49. Accordingly, according to the balancing shaft 220 illustrated in FIG. 51, the vibrational force Fr can also be balanced without generating a large moment.

In this regard, in the above ninth embodiment, the weight for balancing the vibrational force Fr caused by the connecting rod 36 is reflected significantly on the side near the connecting rod 36. However, the feature is not essential for the present invention. Thus, the weight (mr / 2) that is subtracted from the crankshaft 204 to balance the vibrational force Fr can be evenly subtracted from the two eccentric weights 208 of the crankshaft. Similarly, for the balancing shaft 206, the weight (mr / 2) can be evenly subtracted from the entire area of the balancing shaft 206.

Further, in the above ninth embodiment, the configuration for balancing the vibration causing force Fr caused by the connecting rod 36 is included in the balancing device based on the first embodiment, but the present invention is not limited thereto. Thus, it is also possible to include a configuration for balancing the vibration-causing force Fr in a balancing device based on any of the second to seventh embodiments.

Tenth Embodiment

Configuration of Tenth Embodiment

Next, a tenth embodiment of the present invention is explained with reference to FIG. 52-60.

The balancing device 30 (see FIG. 1) of the above first embodiment is configured in such a way that the connection point 38 to the balancing shaft can slide in the radial direction of the balancing shaft 40. Hereinafter, this type is called the “sliding type”. Of the balancing devices disclosed here, the balancing devices in the second embodiment (see FIG. 14), the sixth embodiment (see FIG. 28), the seventh embodiment (see FIG. 35), the eighth embodiment (see FIG. .39) and the ninth embodiment (see FIG. 46) are of a sliding type, in addition to the balancing device of the first embodiment.

Meanwhile, in the balancing device 78 of the third embodiment, the connection point 38 to the balancing shaft is connected to the balancing shaft 80 in such a way that only relative rotation is provided. Later in this document, this type is called the "link type". Of the balancing devices disclosed herein, the balancing devices in the fourth embodiment (see FIG. 19) and the fifth embodiment (see FIG. 25) are of the link type, in addition to the balancing device of the third embodiment.

FIG. 52 is an exploded perspective view of a balancing device 224 of a tenth embodiment of the present invention. The balancing device 224 of the present embodiment has a feature in the aspect of arranging one example of a specific configuration implementing the link type balancing device described above.

The balancing device 224 includes a crankshaft 226. The crankshaft 226 includes a crank pin 228. The crank pin 288 is connected to a piston (not illustrated) of the internal combustion engine through a connecting rod (not illustrated). Eccentric weights 230 of the crankshaft are provided on both sides of the crank pin 228. Additionally, the crankshaft 226 includes a main shaft 232 of the crankshaft as a rotational shaft.

The balancing device 224 includes an eccentric shaft 234 on the side of the crankshaft, which is attached to the main shaft 232 of the crankshaft. The eccentric shaft 234 on the side of the crankshaft contains a through hole 236, which is interconnected for insertion with the main shaft 232 of the crankshaft. A through hole 236 is provided at a position that is eccentric by a fixed value from the center of the eccentric shaft 234 on the crankshaft side. The through hole 236 comprises a positioning groove 238. An eccentric shaft 234 on the crankshaft side is connected to the main shaft 232 of the crankshaft so that the positioning guide 240 on the main shaft 232 of the crankshaft engages with the positioning groove 238. As a result, the eccentric shaft 234 on the side of the crankshaft the shaft is attached to the main shaft 232 of the crankshaft in a state in which relative rotation is not permitted. The eccentric shaft 234 on the crankshaft side contains a plurality of lightening holes 242 at positions that do not interfere with the through hole 236.

The balancing device 224 includes a balancing shaft 244. The balancing shaft 244 includes an eccentric load 246 of the balancing shaft. Additionally, balancing shaft 244 includes an axial shaft 248 of the balancing shaft, which is parallel to the main shaft 232 of the crankshaft. The balancing shaft 244 can rotate with the axial shaft 248 of the balancing shaft as a rotational shaft.

The eccentric shaft 250 on the side of the balancing shaft is connected to the axial shaft 248 of the balancing shaft. The eccentric shaft 250 on the side of the balancing shaft contains a through hole 252, which is interconnected for insertion with the axial shaft 248 of the balancing shaft. A through hole 252 is provided at a position that is eccentric by a fixed value from the center of the eccentric shaft 250 on the side of the balancing shaft. The eccentric shaft 250 on the balancing shaft side is attached to the axial shaft 248 of the balancing shaft in a state in which relative rotation is not permitted, as in the case of the eccentric shaft 234 on the crankshaft side. The eccentric shaft 250 on the side of the balancing shaft further comprises a plurality of relief holes 254 at positions that do not interfere with the through hole 252.

Balancing device 224 includes a connecting rod 256.

FIG. 53 is an exploded perspective view of a connecting rod 256. As illustrated in FIG. 53, connecting rod 256 includes a connecting rod main body 258. The connecting rod main body 258 includes an annular portion 260 on the crankshaft side and an annular portion 262 on the balancing shaft side. The annular portion 260 on the crankshaft side and the annular portion 262 on the balancing shaft side are integrally formed by the connecting portion 264.

After the bearing 266 on the crankshaft side is placed in the annular portion 260 on the crankshaft side, the retaining ring 268 is attached to the annular portion 260 on the crankshaft side. Circlip 268 prevents the lowering of bearing 266 on the crankshaft side. Similarly, after the bearing 270 on the balancing shaft side is placed in the annular portion 262 on the balancing shaft side, the retaining ring 272 is attached to the annular portion 262 on the balancing shaft side. The retaining ring 272 prevents the lowering of the bearing 270 on the side of the balancing shaft. Bearing 266 on the side of the crankshaft and bearing 270 on the side of the balancing shaft are roller bearings comprising a plurality of bearing balls.

The hinge 274 is inserted into the connecting portion 264 of the connecting rod 256. The hinge 274 can act as a roller bearing. The hinge 274 is secured by a nut 276, which is located on the opposite side with the connecting part 264 between the hinge 274 and the nut 276.

Further explanation continues with reference to FIG. 52. The above eccentric shaft 234 on the crankshaft side is inserted into the bearing 266 on the crankshaft side. As a result, the eccentric shaft 234 on the crankshaft side is rotatably held by the connecting rod 256. Similarly, the above eccentric shaft 250 on the balancing shaft side is inserted into the bearing 270 on the balancing shaft side. As a result, the eccentric shaft 250 on the side of the balancing shaft is rotatably held by a connecting rod 256.

The balancing device 224 includes a guide section 278. The guide section 278 has an elongated hole 280 inside. The guide section 278 is fixed in a predetermined position so that the hinge 274 is placed in the elongated hole 280. The oblong hole 280 has a width slightly larger than the diameter of the hinge 274 and directs its movement so that the hinge 274 draws a figure eight path similar to the guide section 8 in the third embodiment.

FIG. 54 is a side cross-sectional view of balancing device 224 of the present embodiment. In FIG. 54, a section of the balancing shaft 244 is drawn in the upper side, and a section of the crankshaft 26 is drawn in the lower side, respectively. In an internal combustion engine, the crankshaft 226 acts as a source of actuation of the oil pump, as well as a source of actuation of the timing chain, which drives the intake and exhaust valves. Here, as one example, a state is illustrated in which an asterisk 282 is provided on a crankshaft main shaft 232, and a timing chain 284 is located on an asterisk 282.

As illustrated in FIG. 54, the eccentric shaft 234 on the crankshaft side is attached to the end of the top of the main shaft 232 of the crankshaft by means of a mounting bolt 286. Hereinafter, the center point of the main shaft 232 of the crankshaft is called “C232” and the center point of the eccentric shaft 234 on the crankshaft the shaft is called "C234". In the present embodiment, an eccentric value A is provided between point C232 and point C234.

As described above, the main shaft 232 of the crankshaft is attached to the eccentric shaft 234 on the side of the crankshaft so that relative rotation does not occur. Therefore, when the main shaft 232 of the crankshaft rotates, the eccentric shaft 234 on the side of the crankshaft rotates together with the main shaft 232 of the crankshaft while maintaining the eccentric value A. At this time, point C234 draws a circular path having a radius A with point C232 as the center .

The eccentric shaft 234 on the crankshaft side is rotatably held in the connecting rod 256. Therefore, the connecting rod 256 can rotate around the eccentric shaft 234 on the crankshaft side with a point C234 as the center of rotation. The eccentric shaft 234 on the crankshaft side is integral with the crankshaft 226. Accordingly, the connecting rod 256 and the crankshaft 226 can relatively rotate with the point C234 as the center. From this point, the center point C234 of the eccentric shaft 234 on the crankshaft side corresponds to the connection point 32 with the crankshaft in the above third embodiment.

When the main shaft 232 of the crankshaft rotates 90 (° CA) from the state illustrated in FIG. 54, point C234 drops to a height identical to the height of point C232. During this rotation, the connecting rod 256 moves down a distance A in FIG. 54. When the main shaft 232 of the crankshaft rotates an additional 90 (° CA), point C234 moves to a point that is downward from point C232 at a distance A. During this rotation, the connecting rod 256 additionally moves to a distance of A, s in order to reach the movable end in the lower side of FIG. 54. Thus, according to the configuration of the present embodiment, when the crankshaft 226 is rotated, the connecting rod 256 can forcibly form a reciprocating movement over a travel distance 2A.

In this regard, in the present embodiment, the crankshaft main shaft 232 has a radius B. Therefore, the distance from the center point C234 of the eccentric shaft 234 on the crankshaft side to the outer circumference of the crankshaft main shaft 232 is a maximum of “A + B”. In the present embodiment, the diameter of the eccentric shaft 234 on the crankshaft side is set to a sufficiently large value so that the distance A + B is sufficiently within the radius (D / 2) of the eccentric shaft 234 on the crankshaft side. Therefore, according to the present embodiment, the through hole 236 (see FIG. 52), which can accommodate the entire diameter of the main shaft 232 of the crankshaft, can be provided in the outer diameter of the eccentric shaft 234 on the side of the crankshaft without losing the required strength.

As illustrated in the upper side in FIG. 54, the eccentric shaft 250 on the balancing shaft side is connected to the end of the top of the axial shaft 248 of the balancing shaft by means of a set screw 288. Hereinafter, the center point of the axial shaft 248 of the balancing shaft is called “C248” and the central point of the eccentric shaft 250 on the side of the balancing shaft the shaft is called "C250". In the present embodiment, an eccentric value A is provided between point C248 and point C250, similarly to the side of the main shaft 232 of the crankshaft.

When the crankshaft 226 rotates so that the center point C234 of the eccentric shaft 234 on the crankshaft side rotates from the state illustrated in FIG. 54, toward the front of the drawing with the point C232 as the center, the guide section 278 functions similarly to the guide section 82 in the third embodiment. In particular, the guide section 278 does not allow the hinge 274 to move toward the front side of the drawing at this time and restricts the movement of the hinge 274 to the back side of the drawing. As a result, a movement that biases the eccentric shaft 250 on the side of the balancing shaft to the rear side of the drawing occurs in the connecting rod 256.

The movement of the eccentric shaft 250 on the side of the balancing shaft is limited by the connecting rod 256 and the axial shaft 248 of the balancing shaft. On the other hand, the eccentric shaft 250 on the side of the balancing shaft can rotate in the connecting rod 256. According to the limitation of the connecting rod 256, the eccentric shaft 250 on the side of the balancing shaft must be shifted to the lower side of the drawing to move from the state illustrated in FIG. 54, to the rear side of the drawing, simultaneously. Additionally, the point C248 is fixed by the axial shaft 248 of the balancing shaft, and therefore, the above-described displacement should occur while the distance between the point C250 and the point C248 is maintained at an eccentric value A.

As the rotation of the crankshaft 226 continues from the state illustrated in FIG. 54, under the restrictive condition as indicated above, the eccentric shaft 250 on the side of the balancing shaft exhibits such a motion that it revolves around the axial shaft 248 of the balancing shaft so that its center point C250 draws a circular orbit having a radius A with point C248 as center. As described above, the eccentric shaft 250 on the side of the balancing shaft is attached to the axial shaft 248 of the balancing shaft so that it does not rotate relatively. Therefore, the rotation of the eccentric shaft 250 on the side of the balancing shaft is directly transmitted to the axial shaft 248 of the balancing shaft.

In the configuration of the present embodiment, the connecting rod 256 can rotate relative to the eccentric shaft 250 on the side of the balancing shaft with an axial line passing through point C250 as the axis of rotation. The eccentric shaft 250 on the side of the balancing shaft is integral with the balancing shaft 244. Accordingly, the connecting rod 256 and the balancing shaft 244 can relatively rotate with the point C250 as the center. In this regard, the center point C250 of the eccentric shaft 250 on the balancing shaft side corresponds to the connection point 38 with the balancing shaft in the link type balancing device (see FIG. 15).

While the eccentric shaft 250 on the side of the balancing shaft revolves around the axial shaft 248 of the balancing shaft once, the point C250 exhibits a reciprocating movement over a travel distance 2A. During this movement, the connecting rod 256 likewise exhibits a reciprocating movement over a stroke distance 2A. The reciprocating movement is performed synchronously with the reciprocating movement that occurs on the side of the main shaft 232 of the crankshaft. As a result, according to the configuration of the present embodiment, the operation and function explained in the third embodiment can be realized.

In this regard, in the present embodiment, the balancing shaft axial shaft 248 has a radius E. Therefore, the distance from the center point C250 of the eccentric shaft 250 on the balancing shaft side to the outer circumference of the balancing shaft axial shaft 248 is a maximum of “A + E”. In the present embodiment, the diameter F of the eccentric shaft 250 on the side of the balancing shaft is set to a sufficiently large value, so that the distance A + E is located sufficiently within the radius (F / 2) of the eccentric shaft 250 on the side of the balancing shaft. Therefore, according to the present embodiment, the through hole 252 (see FIG. 52), which can accommodate the entire diameter of the axial shaft 248 of the balancing shaft, can be provided in the outer diameter of the eccentric shaft 250 on the side of the balancing shaft without losing the required strength.

Method for machining an eccentric shaft on the side of the crankshaft and an eccentric shaft on the side of the balancing shaft

FIG. 55 is a diagram for explaining a method for forming an eccentric shaft 234 on a side of a crankshaft and an eccentric shaft 250 on a side of a balancing shaft in the present embodiment by means of a general hole machining. As described above, in the balancing device 224 of the present embodiment, the eccentric value of the main shaft 232 of the crankshaft and the eccentric shaft 234 on the side of the crankshaft and the eccentric value of the axial shaft 248 of the balancing shaft and the eccentric shaft 250 on the side of the balancing shaft are set to the same value as “A” .

As described for the third embodiment, in the link type balancing device, the rotation radius r1 of the connection point with the crankshaft and the rotation radius r2 of the connection point with the balancing shaft should have identical values. In the present embodiment, the connection point with the crankshaft is C234, and its rotation radius r1 corresponds to the eccentric value A. Similarly, the connection point to the balancing shaft is C250, and its rotation radius r2 corresponds to the eccentric value A. Accordingly, in the present embodiment of implementation, the eccentric value A on the side of the main shaft 232 of the crankshaft and the eccentric value A on the side of the axial shaft 248 of the balancing shaft must exactly match.

FIG. 55 specifically illustrates the state in which the eccentric shaft 250 on the side of the balancing shaft and the eccentric shaft 234 on the side of the crankshaft are defined in an identical jig 290. The jig 290 has grooves 292 and 294 with two ledges with an axial line illustrated by assigning C234, C250 in FIG. 55, as a center. The groove 292 is formed with a circular shape having a diameter equal to the diameter of the eccentric shaft 234 on the side of the crankshaft. Additionally, the groove 294 is formed with a circular shape having a diameter equal to the diameter of the eccentric shaft 250 on the side of the balancing shaft. These grooves 292 and 294 are machined so that they are coaxial with each other.

In the present embodiment, the eccentric shaft 250 on the side of the balancing shaft in the stage at which the machining of the external shape and the relief holes 242 is completed is defined in the groove 294 of the jig 290 first. Then, the eccentric shaft 234 on the side of the crankshaft at the stage where the machining of the external shape and the relief holes 254 is completed is defined in the groove 292 of the clamping device 290. After that, the through hole 252 of the eccentric shaft 250 on the side of the balancing shaft and the through hole 236 of the eccentric shaft 234 on the side of the crankshaft are sequentially provided by a method of general hole machining. According to such a method, the eccentric value A of the eccentric shaft 234 on the crankshaft side and the eccentric value A of the eccentric shaft 250 on the side of the balancing shaft can exactly match.

Effect of lightening holes

FIG. 56 is a front view of an internal combustion engine including a balancing device 224 of the present embodiment. More specifically, the left side in FIG. 56 expresses the state in which the piston 12 is located at top dead center. Additionally, the right side in FIG. 56 expresses the state in which the piston 12 is located at bottom dead center.

The balancing device 224 of the present embodiment has such a configuration that the connecting rod 256 reaches the movable end in the upper right side of FIG. 56, when the piston 12 reaches top dead center and the connecting rod 256 reaches the moving end in the lower left side of FIG. 56 when the piston 12 reaches bottom dead center. Thus, the balancing device 224 of the present embodiment has a configuration that causes the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft to work in almost the opposite direction relative to the connecting rod 256.

In a balancing device with such a configuration, the weight mr of the connecting rod 256 must be added to the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft in order to balance the vibrational force Fr caused by the connecting rod 256. Therefore, from the point of view of reducing the weight of the internal combustion engine than the lighter the connecting rod 256, the more desirable it is. In this sense, the lighter the eccentric shaft 234 on the side of the crankshaft and the eccentric shaft 250 on the side of the balancing shaft, which move as a unit with the connecting rod 256, the more desirable.

Additionally, in the balancing device 224 of the present embodiment, the wide part (the part in which the lightening holes 242 are formed) of the eccentric shaft 234 on the crankshaft side and the wide part (the part in which the lightening holes 254 are formed) of the eccentric shaft 250 on the balancing shaft side reach to the movable end in the upper right side of FIG. 56 in synchronization with the time at which the connecting rod 256 reaches the movable end in the same direction in FIG. 56. Therefore, if the wide parts have large weights, a large vibrational force is due to the centrifugal force of the eccentric shaft 234 itself on the crankshaft side and the centrifugal force of the eccentric shaft 250 itself on the side of the balancing shaft. Thus, weight reduction is particularly desirable with respect to a wide portion of the eccentric shaft 234 on the crankshaft side and a wide portion of the eccentric shaft 250 on the balancing shaft side.

As illustrated in FIG. 56, in the present embodiment, the relief holes 242 and 254 are provided in wide parts. According to the relief holes 242 and 254, the weights of the above wide parts can be significantly reduced. As a result, the vibration causing force caused by the centrifugal force of the eccentric shaft 234 on the crankshaft side and the centrifugal force of the eccentric shaft 250 on the balancing shaft side is significantly reduced. Additionally, the vibration causing force caused by the weight of the connecting rod 256 is also significantly reduced. Therefore, according to the configuration of the present embodiment, the weight reduction of the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft is achieved. This leads to a decrease in the weight of the internal combustion engine.

An example of a modification of the tenth embodiment of the invention

FIG. 57 is a side cross-sectional view of an example of a modification of the balancing device 224 of the present embodiment. The balancing device 224 of the present embodiment is designed so that the diameter of the main shaft 232 of the crankshaft is sufficiently contained within the external shape of the eccentric shaft 234 on the side of the crankshaft, as described above. Therefore, the crankshaft main shaft 232 can penetrate through the eccentric shaft 234 on the crankshaft side while maintaining a circular shape.

Additionally, in the balancing device 224 of the present embodiment, the hinge 274 and the guide section 278 are located on the side of the crankshaft 226 and the balancing shaft 244, due to which the surface side (left side in Fig. 57) of the connecting rod 256 can become almost flat. Therefore, according to the balancing device 224, the main shaft 232 of the crankshaft protrudes from the connecting rod 256 for a considerable length and can be used to drive various devices. FIG. 57 illustrates an example in which the drive shaft 296 of the oil pump is inserted into the main shaft 232 of the crankshaft, in addition to 282 for actuating the intake and exhaust valves. Thus, according to the balancing device 224 of the present embodiment, the size reduction of the internal combustion engine can be stimulated.

FIG. 58 is a diagram for explaining a configuration of an example of a modification of an eccentric shaft on a balancing shaft side that can be used in the present embodiment. The eccentric shaft 298 on the balancing shaft side, illustrated in FIG. 58 has a diameter G, i.e. radius (G / 2). The radius (G / 2) sometimes cannot keep from becoming less than the sum of the eccentric value A and the radius E of the axial shaft 248 of the balancing shaft due to various restrictions. The configuration illustrated in FIG. 58 illustrates an example solution in such a case. In the configuration, the eccentric shaft 298 on the balancing shaft side has a smaller through hole than the axial shaft section 248 of the balancing shaft. On the other hand, the tip end of the axial shaft 248 of the balancing shaft is machined into a mold that enables insertion of the tip end into the through hole. The machined end of the top of the axial shaft 248 of the balancing shaft is mounted in a state in which the end of the top is inserted into the through hole of the eccentric shaft 298 on the side of the balancing shaft. According to such a configuration, even if a sufficiently large space cannot be provided for the eccentric shaft 298 on the side of the balancing shaft, a balancing device using a connecting rod 256 can be implemented.

FIG. 59 is a diagram for explaining a configuration of an example of a modification of an eccentric shaft on the crankshaft side, which is applicable in the present embodiment. The eccentric shaft 300 on the crankshaft side illustrated in FIG. 59 has a diameter H, i.e. radius (H / 2). The radius (H / 2) sometimes cannot keep from becoming less than the sum of the eccentric value A and radius B of the main shaft 232 of the crankshaft due to various restrictions. The configuration illustrated in FIG. 59, shows one example of a solution in such a case. In the configuration, the eccentric shaft 300 on the crankshaft side has a smaller through hole than the section of the main shaft 232 of the crankshaft. On the other hand, the tip end of the main shaft 232 of the crankshaft is machined into a mold that enables insertion of the tip end into the through hole. The machined end of the top of the main shaft 232 of the crankshaft has a mounting bolt 288 attached thereto in a state in which the end of the top is inserted into the through hole of the eccentric shaft 300 on the side of the crankshaft. According to such a configuration, even if a sufficiently large space cannot be provided for the eccentric shaft 300 on the crankshaft side, a balancing device using a connecting rod 256 can be implemented.

FIG. 60 is a diagram illustrating an example of a modification of the balancing device of the present embodiment in front view. In particular, the left side in FIG. 60 shows a state in which the piston 12 is located at top dead center in a modification example. Additionally, the right side in FIG. 60 shows a state in which the piston 12 is located at bottom dead center in a modification example.

The balancing device 302 illustrated in FIG. 60 has a configuration in which the connecting rod 256 reaches the movable end in the lower left side of FIG. 60, when the piston 12 reaches top dead center and the connecting rod 256 reaches the moving end in the upper right side in FIG. 60 when the piston 12 reaches bottom dead center. Thus, in the balancing device 302, the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft operate in a direction almost identical to the direction of the connecting rod 256.

In a balancing device with this configuration, the vibration causing force Fr caused by the connecting rod 256 can be used as the force that balances the vibration causing force Fp + Fc caused by the piston 12 and the connecting rod 14. As the vibration causing force Fr becomes more, the weights to be provided to the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft can be set smaller.

An eccentric shaft 304 on the side of the crankshaft illustrated in FIG. 60 does not have a lightening hole, unlike the eccentric shaft 234 on the crankshaft side in the tenth embodiment. Similarly, the eccentric shaft 306 on the balancing shaft side illustrated in FIG. 60 does not have a lightening hole, unlike the eccentric shaft 250 on the balancing shaft side in the tenth embodiment. According to the eccentric shaft 304 on the side of the crankshaft and the eccentric shaft 306 on the side of the balancing shaft, as described above, a large vibrational force can be generated by the centrifugal forces themselves, and the connecting rod 256 can forcefully generate a large vibrational force Fr. Therefore, the configuration illustrated in FIG. 60 may help reduce the weight of the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft.

Additionally, in the above tenth embodiment, the bearing 266 on the side of the crankshaft and the bearing on the side of the balancing shaft, which are contained in the connecting rod 256, are configured by roller bearings. However, these bearings are not limited to roller bearings. For example, these bearings may be plain bearings using lubricating oil. The aspect likewise applies to the eleventh embodiment, which is described below.

Eleventh Embodiment

Next, an eleventh embodiment of the present invention is explained with reference to FIG. 61-66.

FIG. 61 is an exploded perspective view of a balancing device 308 of an eleventh embodiment of the present invention. The balancing device 308 of the present embodiment has a feature in the aspect of arranging one example of a specific configuration that implements a sliding type balancing device illustrated in the first embodiment (see FIG. 1), and the like. Further in this document, in FIG. 61-66, identical or common elements with elements in the above tenth embodiment, identical reference numbers are assigned, and their explanation is omitted or simplified.

In balancing device 308, an eccentric shaft 234 on the side of the crankshaft is attached to the main shaft 232 of the crankshaft, as in the case of the tenth embodiment. The eccentric shaft 310 on the side of the balancing shaft is attached to the axial shaft 248 of the balancing shaft. The eccentric shaft 310 on the side of the balancing shaft has an elongated hole 312 on the surface opposite the surface that is attached to the axial shaft 248 of the balancing shaft. The outer shape of the eccentric shaft 310 on the side of the balancing shaft is cylindrical, and an elongated hole 312 is provided along the radial direction of the circle.

The balancing device 308 includes a connecting rod 314, which connects the eccentric shaft 234 on the side of the crankshaft and the eccentric shaft 310 on the side of the balancing shaft.

FIG. 62 is an exploded perspective view of a connecting rod 314. As illustrated in FIG. 62, the connecting rod 314 has a connecting rod main body 316. The connecting rod main body 316 has a column portion 318 that extends from the annular portion 260 on the crankshaft side. The hinge 322 of the balancing shaft is attached to the end of the columnar part 318 by means of a nut 320. Additionally, in the columnar part 318, the hinge 274 is attached to the vicinity of the boundary relative to the annular part 260 on the side of the crankshaft by the nut 276. The hinge 322 of the balancing shaft acts as a roller bearing, similar to hinge 274.

Further explanation continues with reference to FIG. 61. The above eccentric shaft 234 on the crankshaft side is inserted into the bearing 266 on the crankshaft side of the connecting rod 314, as in the case of the tenth embodiment. On the other hand, the eccentric shaft 310 on the side of the balancing shaft is connected to the connecting rod 314 by placing the hinge 322 of the balancing shaft in the elongated hole 312.

The balancing device 308 includes a guide section 324. The guide section 324 has an elongated hole 326 inside. The oblong hole 326 has a width corresponding to the diameter of the hinge 274, and adjusts the movement of the hinge 274 to a rectilinear motion, similar to the guide section 54 in the first embodiment.

FIG. 63 is a side cross-sectional view of a balancing device 308 of the present embodiment. In FIG. 63, the balancing shaft section 244 is drawn in the upper side, and the crankshaft section 226 is drawn in the lower side, respectively. The crankshaft 226 comprises an asterisk 282, which drives the timing chain 284, as in the case of the tenth embodiment.

In the configuration illustrated in FIG. 63, the eccentric shaft 234 on the crankshaft side operates as in the case of the tenth embodiment. In particular, in the present embodiment, the center point C234 of the eccentric shaft 234 on the crankshaft side also corresponds to the connection point 32 with the crankshaft of the balancing device. When the crankshaft 226 rotates, its connection point 32 with the crankshaft (C234) draws a circular path around the center point C232 of the main shaft of the crankshaft while maintaining the eccentric value A. As a result, the connecting rod 314 performs a reciprocating movement at a distance of 2A.

It should be noted that in the present embodiment, the radius (D / 2) of the eccentric shaft 234 on the crankshaft side is set to a sufficiently large value with respect to the sum of the eccentric value A and the radius B of the main shaft 232 of the crankshaft. Therefore, in the construction of the present embodiment, it is also possible to apply a design suitable for downsizing, as illustrated in FIG. 57.

As illustrated in the upper side in FIG. 63, the eccentric shaft 310 on the side of the balancing shaft is attached to the end of the top of the axial shaft 248 of the balancing shaft by means of a mounting bolt 328. The hinge 322 of the balancing shaft of the connecting rod 314 is placed in the elongated hole 312 of the eccentric shaft 310 on the side of the balancing shaft. Hereinafter, the center point of the axial shaft 248 of the balancing shaft is called “C248”, and the center point of the hinge 322 of the balancing shaft is called “C322”.

In the configuration of the present embodiment, the connecting rod 314 can rotate relative to the eccentric shaft 310 on the side of the balancing shaft with an axial line passing through point C322 as the axis of rotation. The eccentric shaft 310 on the side of the balancing shaft is integral with the balancing shaft 244. Accordingly, the connecting rod 314 and the balancing shaft 244 can relatively rotate with the point C322 as the center. From this point, the center point C322 of the balancing shaft hinge 322 corresponds to the connection point 38 with the balancing shaft in the sliding type balancing device (see FIG. 2).

As explained with reference to FIG. 2, in the balancing device 30 of the sliding type, the distance between the connection point 38 with the balancing shaft and the center point of the balancing shaft 40 changes in accordance with the operation of the connecting rod 36. The balancing device 308 of the present embodiment also has a sliding type, and therefore, the distance between the connection point 38 with balancing shaft and center point of balancing shaft 244, i.e. the distance between the center point C322 of the balancing shaft hinge 322 and the center point C248 of the axial shaft 248 of the balancing shaft should be variable.

FIG. 64 is a diagram showing a positional relationship between the center point C248 of the balancing shaft axial shaft 248 and the center point C322 of the balancing shaft hinge 322 when viewed along the LXIV arrows in FIG. 63. As illustrated in FIG. 64, according to the configuration of the present embodiment, the eccentric value I C322-C248 is changed by smoothly moving the balancing shaft hinge 322 along the oblong hole 312. Therefore, according to the configuration of the eccentric shaft 310 on the side of the balancing shaft and the connecting rod 314 in the present embodiment, the requirement can be satisfied. necessary for the operation of the balancing device of the sliding type.

In addition, the operation of the balancing device 308 is described below with reference to FIG. 63. When the crankshaft 226 rotates from the state illustrated in FIG. 63, so that the center point C234 of the eccentric shaft 234 on the crankshaft side rotates to the front side of the drawing with the point C232 as the center, the guide section 324 functions similarly to the guide section 54 in the first embodiment. In particular, at this time, the guide section 324 does not allow the hinge 274 to move toward the front side of the drawing and restricts the movement of the hinge 274 to only the movement toward the underside of the drawing. As a result, a movement occurs in the connecting rod 314 that biases the hinge 322 of the balancing shaft to the rear side and the lower side of the drawing.

The movement of the hinge 322 of the balancing shaft imparts rotation in the opposite direction relative to the rotation of the crankshaft 226 for the eccentric shaft 310 on the side of the balancing shaft. Since the eccentric shaft 310 on the side of the balancing shaft is attached to the axial shaft 248 of the balancing shaft, rotational movement occurs in the axial shaft 248 of the balancing shaft in synchronization with the offset of the hinge 322 of the balancing shaft. Then, as the crankshaft 226 continues to rotate, the center point C322 of the balancing shaft hinge 322 continues to rotate around C248, and the balancing shaft 244 rotates continuously. As a result, vertical movement with a distance of 2A as a stroke length occurs in the crank 314.

Effect of lightening holes

FIG. 65 is a diagram showing an internal combustion engine including a balancing device 308 of the present embodiment in front view. More specifically, the left side in FIG. 65 shows a state in which the piston 12 is located at top dead center. Additionally, the right side in FIG. 65 shows a state in which the piston 12 is located at bottom dead center.

The balancing device 308 of the present embodiment has the configuration of the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft in almost the opposite direction relative to the connecting rod 314, as in the case of the tenth embodiment. In a balancing device with this configuration, the lighter the connecting rod 314, the more desirable this is, in terms of reducing the weight of the internal combustion engine. In the present embodiment, the relief holes 242 are provided in a wide part of the eccentric shaft 234 on the crankshaft side, as in the case of the tenth embodiment. Therefore, according to the configuration of the present embodiment, the weight of the internal combustion engine can also be reduced, as in the case of the tenth embodiment.

Modification of the Eleventh Embodiment

As described above, in the eleventh embodiment, the eccentric shaft 234 on the crankshaft side is provided with a sufficiently large radius (D / 2) compared to the sum of the eccentric value A and radius B of the main shaft 232 of the crankshaft. However, there may be a case where the radius (D / 2) of the main shaft 232 of the crankshaft cannot keep from becoming less than A + B due to various restrictions. In the balancing device 308 of the present embodiment, the configuration, as illustrated in FIG. 59 may be adapted in such a case as in the case of the tenth embodiment.

FIG. 66 is a front view of an example of a modification of the balancing device of the present embodiment. In particular, the left side in FIG. 66 shows a state in which the piston 12 is located at top dead center in a modification example. Additionally, the right side in FIG. 66 shows a state in which the piston 12 is located at bottom dead center in a modification example. In the balancing device 330 illustrated in FIG. 66, the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft operate in a direction almost identical to the direction of the connecting rod 314, in contrast to the case of the eleventh embodiment.

In balancing device 330 with this configuration, the weight of the connecting rod 314 can be used to balance the vibrational force Fr caused by the piston 12 and the connecting rod 14. Therefore, in the balancing device 330, the weights to be provided to the eccentric load 230 of the crankshaft and the eccentric the load 246 of the balancing shaft can be reduced by leaving a lot of weight in the connecting rod 314. Thus, in the configuration illustrated in FIG. 66, a relief hole is not provided in the eccentric shaft on the crankshaft side, in contrast to the case of the eleventh embodiment. According to this configuration, a reduction in the weight of the eccentric load 230 of the crankshaft and the eccentric load 246 of the balancing shaft can be achieved.

Claims (70)

1. A balancing device for an internal combustion engine, comprising: a crankshaft that rotates with the main shaft of the crankshaft as a rotational shaft; and a balancing shaft that rotates with the axial shaft of the balancing shaft parallel to the main shaft of the crankshaft as a rotational shaft, wherein the crankshaft includes an eccentric load of the crankshaft, which sets the center of gravity of the crankshaft eccentric relative to the center of the main shaft of the crankshaft, and the balancing shaft includes an eccentric load of the balancing shaft, which sets the center of gravity of the balancing shaft to be eccentric relative to the center of the axial shaft of the balancing shaft, moreover, the balancing device further comprises: a connecting rod, which connects the connection point with the crankshaft located on the crankshaft in a position deviated from the center of the main shaft of the crankshaft, and the connection point with the balancing shaft located on the balancing shaft in a position deviated from the center of the axial shaft of the balancing shaft; a mechanism for connecting to the crankshaft, which provides relative rotation of the crankshaft and connecting rod with a connection point with the crankshaft as the center of rotation; a mechanism for connecting to the balancing shaft, which provides relative rotation of the balancing shaft and connecting rod with a connection point with the balancing shaft as the center of rotation; and a guide section that guides the movement of the connecting rod in such a way that the balancing shaft rotates in the opposite direction relative to the direction of rotation of the crankshaft. 2. A balancing device for an internal combustion engine according to claim 1, further comprising: a mechanism for adjusting the connection points, which provides at least one of the connection point with the crankshaft and the connection point with the balancing shaft the possibility of shifting in the direction of the radius of rotation of the specified at least one of the connection point with the crankshaft and the connection point with the balancing shaft; and the sliding part located at one point of the connecting rod, wherein the guide section controls the movement of the sliding part to a rectilinear movement in the direction from the side of the main shaft of the crankshaft to the side of the axial shaft of the balancing shaft and the rectilinear movement in the opposite direction with respect to the specified direction. 3. A balancing device for an internal combustion engine according to claim 1, further comprising: a mechanism for adjusting the connection points, which provides at least one of the connection point with the crankshaft and the connection point with the balancing shaft the possibility of shifting in the direction of the radius of rotation of the specified at least one of the connection point with the crankshaft and the connection point with the balancing shaft; however, the guide section is made to rotate in a plane identical to the connecting plane of the connecting rod with a position superimposed on the connecting rod as the center, and holding the connecting rod with the possibility of sliding in the direction of the center line of the connecting rod. 4. A balancing device for an internal combustion engine according to claim 1, further comprising: a restrictive part located at a midpoint between the connection point with the crankshaft and the connection point with the connecting rod balancing shaft, the distance between the connection point with the crankshaft and the connection point with the balancing shaft is equal to the distance between the main shaft of the crankshaft and the axial shaft of the balancing shaft, and the distance between the center of the main shaft of the crankshaft and the connection point with the crankshaft is equal to the distance between the center of the axial shaft of the balancing shaft and the connection point with the balancing shaft, moreover, the guide section includes a guide on the side of the balancing shaft, which prevents the restrictive part from shifting in the direction of rotation identical to the direction of rotation of the connection point with the crankshaft in a position in which the restrictive part is closest to the axial shaft of the balancing shaft, and the guide on the side crankshaft, which prevents the restrictive part from shifting in the direction of rotation identical to the direction of rotation of the connection point with the crankshaft waste in a position in which the restrictive part approaches the most to the main shaft of the crankshaft. 5. A balancing device for an internal combustion engine according to claim 2, in which: the crankshaft is used by the offset crank method, in which the center of the main shaft of the crankshaft is set in a position that is offset by a fixed value from the center line of the reciprocating motion of the piston, and the balancing shaft and the guide section are arranged so that at least one of the center of the main shaft of the crankshaft and the center of the axial shaft of the balancing shaft is placed in a position that is shifted by a fixed value from the axial line of rectilinear motion. 6. The balancing device for the internal combustion engine according to claim 5, in which the center line "crankshaft - balancing shaft" connecting the center of the main shaft of the crankshaft and the center of the axial shaft of the balancing shaft is shifted by a fixed value from the axial line of rectilinear movement. 7. A balancing device for an internal combustion engine according to claim 5, in which: the center of the main shaft of the crankshaft is located on the axial line of rectilinear motion, and the center of the axial shaft of the balancing shaft is offset by a fixed value from the axial line of the rectilinear motion. 8. The balancing device for an internal combustion engine according to claim 5, in which: the center of the axial shaft of the balancing shaft is located on the axial line of rectilinear motion, and the center of the main shaft of the crankshaft is offset by a fixed value from the center line of the rectilinear motion. 9. A balancing device for an internal combustion engine according to claim 5, in which: the center of the axial shaft of the balancing shaft is shifted by a fixed value to one side of the axial line of the rectilinear movement, and the center of the main shaft of the crankshaft is shifted by a fixed value to the other side of the axial line of rectilinear motion. 10. The balancing device for an internal combustion engine according to claim 3, in which: the crankshaft is used by the offset crank method, in which the center of the main shaft of the crankshaft is set in a position that is shifted by a fixed value from the reciprocating motion of the piston, and the center of rotation of the guide section is shifted by a fixed value from the center line "crankshaft - balancing shaft" connecting the center of the main shaft of the crankshaft and the center of the axial shaft of the balancing shaft. 11. The balancing device for an internal combustion engine according to claim 1, in which: the connection point with the crankshaft is located on the side identical to the center of gravity of the eccentric load of the crankshaft relative to the center of the main shaft of the crankshaft, and the connection point with the balancing shaft is located on the side identical to the center of gravity of the eccentric load of the balancing shaft relative to the center of the axial shaft of the balancing shaft. 12. The balancing device for an internal combustion engine according to claim 1, in which: the connection point with the crankshaft is located on the opposite side relative to the center of gravity of the eccentric load of the crankshaft relative to the center of the main shaft of the crankshaft, and the connection point with the balancing shaft is located on the opposite side relative to the center of gravity of the eccentric load of the balancing shaft relative to the center of the axial shaft of the balancing shaft. 13. The balancing device for an internal combustion engine according to claim 1, further comprising a torque application mechanism that applies torque in the opposite direction to the balancing shaft relative to the direction of rotation of the crankshaft. 14. The balancing device for an internal combustion engine according to claim 13, in which: the torque application mechanism includes a cam, which is located on the balancing shaft, and a spring element, which is compressed by pressing by the cam, moreover, the cam is capable of pressing the spring element during the movement of the connecting rod to the side of the axial shaft of the balancing shaft with the rotation of the balancing shaft and the adoption of torque in the opposite direction from the spring element in the position in which the axial line of the connecting rod is superimposed on the axial shaft of the balancing shaft. 15. The balancing device for an internal combustion engine according to claim 1, mounted on a single-cylinder or four-stroke two-cylinder internal combustion engine. 16. The balancing device for an internal combustion engine according to claim 1, in which: the connecting rod is inclined relative to the axial line of the reciprocating motion of the piston at the top dead center and bottom dead center of the internal combustion engine, moreover, the eccentric load of the crankshaft has a center of gravity in the region that is on the opposite side relative to the connection point with the crankshaft, with the axial line of the crankshaft, which passes through the center of the main shaft of the crankshaft and is parallel with the axial line of the piston between them, in a situation top dead center the eccentric load of the balancing shaft has a center of gravity in the region that is on the opposite side relative to the connection point with the balancing shaft, with the axial line of the balancing shaft, which passes through the center of the axial shaft of the balancing shaft and is parallel to the axial line of the piston between them, in a situation at top dead center. 17. The balancing device for an internal combustion engine according to claim 16, in which: the eccentric load of the crankshaft has a center of gravity and weight with an absolute value that balances the resulting force from the vibration-induced force caused by the connecting rod of the internal combustion engine, the part that causes the vibration caused by the piston of the internal combustion engine, and the part that causes the vibration caused by the connecting rod , moreover, the eccentric load of the balancing shaft has a center of gravity and weight with an absolute value that balances the remainder of the vibration causing force caused by the piston of the internal combustion engine and the rest of the vibration causing force caused by the connecting rod, wherein said parts of the vibration causing force and said remaining parts of the vibration causing force are equal. 18. The balancing device for an internal combustion engine according to claim 17, in which: the balancing shaft is connected to the connecting rod at one end of the balancing shaft, and From the weight of the eccentric load of the balancing shaft, the weight to balance the remaining part of the vibration-induced force caused by the connecting rod is reflected more significantly in the vicinity of one end compared to the vicinity of the other end of the balancing shaft. 19. A balancing device for an internal combustion engine according to claim 1, in which: the connecting rod has a bearing on the crankshaft side on the crankshaft side, the crankshaft coupling mechanism has an eccentric shaft on the crankshaft side, which is rotatably held by a bearing on the crankshaft side, an eccentric shaft on the side of the crankshaft is attached to the main shaft of the crankshaft so that the eccentric point of the crankshaft, which deviates by a fixed value from its center, coincides with the center of the main shaft of the crankshaft, and the center of the eccentric shaft on the crankshaft side configures the connection point with the crankshaft. 20. The balancing device for an internal combustion engine according to claim 1, in which: the connecting rod has a bearing on the side of the balancing shaft on the side of the balancing shaft, the balancing shaft connection mechanism has an eccentric shaft on the balancing shaft side, which is rotatably held by a bearing on the balancing shaft side, an eccentric shaft on the side of the balancing shaft is attached to the axial shaft of the balancing shaft so that the eccentric point of the balancing shaft, which deviates by a fixed value from its center, coincides with the center of the axial shaft of the balancing shaft, and the center of the eccentric shaft on the balancing shaft side configures the connection point with the balancing shaft.

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