JP5189070B2 - Actuator and variable valve operating apparatus for internal combustion engine to which this actuator is applied - Google Patents

Description

  The present invention relates to a variable valve operating apparatus that variably controls operation characteristics such as a valve lift amount and an operation angle of an intake valve and an exhaust valve of an internal combustion engine, for example, and more particularly to an improved technique for an actuator that drives the variable mechanism.

  For example, various actuators for variable valve operating devices that change the lift amount of an intake valve according to the operating state of a conventional internal combustion engine are provided, and one of them is described in Patent Document 1 below. Are known.

  In brief, by rotating the ball screw shaft by the rotational torque of the electric motor and moving the moving nut in the axial direction, the rotational position of the control shaft is changed via the link mechanism, and thereby the intake valve The lift amount is made variable according to the engine operating state. The moving nut is biased to an intermediate position in the moving direction by the spring force of a pair of bias springs.

  As a result, even if the drive mechanism fails, the moving nut is held at the intermediate position, so that startability during cold operation is ensured.

JP 2008-208780 A

  In the conventional actuator, since the lead angle of the screw formed on the outer periphery of the ball screw shaft is small, the moving nut can be directly secured by the spring force of the bias spring when the internal combustion engine is stopped. It was difficult to move it in the axial direction.

  The present invention has been devised in view of the actual state of the conventional actuator, and the invention according to claim 1 controls the rotation of the output shaft in accordance with the state of the internal combustion engine so that the operating force is applied to the control shaft. An electric motor to be applied, a reversible transmission shaft having a reversible characteristic to which a rotational force is transmitted from the output shaft, and a biasing means for imparting a biasing force to the reversible transmission shaft are provided.

The invention according to claim 2 is an actuator of a variable valve operating apparatus that varies the operating characteristic of the engine valve by moving the control shaft by the power of the electric motor in accordance with the state of the internal combustion engine. It has a reversible transmission shaft that is rotationally driven by a motor and applied a biasing force in one direction, and the transmission efficiency from the control shaft to the electric motor is smaller than the transmission efficiency from the electric motor to the control shaft, The transmission efficiency from the electric motor to the reversible transmission shaft is equal to the transmission efficiency from the reversible transmission shaft to the electric motor.

  According to a third aspect of the present invention, there is provided a drive shaft to which a rotational force is transmitted from a crankshaft and provided with a drive cam, a transmission mechanism for transmitting the power by converting the rotational force of the drive shaft into a swinging force, A swing cam for operating the engine valve by performing a swing motion by transmitting the swing force of the transmission mechanism, and a movement according to the state of the internal combustion engine, thereby changing the transmission path of the transmission mechanism to change the swing mechanism. A variable mechanism including a control shaft that changes a swing state of the dynamic cam and varies a lift amount of the engine valve, an electric motor that controls the rotation of the output shaft in accordance with the state of the internal combustion engine, and the output shaft And a reversible transmission shaft having a reversible characteristic and an urging means for applying a urging force to the reversible transmission shaft.

  According to the present invention, since the rotational force of the reversible transmission shaft urged by the urging means is directly transmitted to the output shaft, the output shaft can be easily rotated, for example, when the engine is stopped. Thus, the control shaft can be easily moved to a desired movement position.

It is a principal part perspective view which shows the variable valve apparatus to which the actuator of 1st Embodiment of this invention was applied. It is a side view which shows a partial cross section of the actuator of this embodiment. It is operation | movement explanatory drawing which shows the intermediate movement position of the ball nut of the actuator. It is operation | movement explanatory drawing which shows a movement position to the maximum right direction of the ball nut of the actuator. 1A is a view as viewed in the direction of the arrow A in FIG. 1 showing the valve closing action during the minimum lift control in the variable valve operating apparatus, and B is a view taken in the direction of the arrow A in FIG. 1A is a view as viewed from an arrow A in FIG. 1 showing a valve closing action at the time of maximum lift control in the variable valve apparatus, and B is a view as seen from an arrow A of FIG. 1 showing a valve opening action at the time of maximum lift control. It is a valve lift characteristic view of each intake valve by the variable valve operating apparatus of this embodiment. It is a side view which shows the actuator of 2nd Embodiment in partial cross section. It is a side view which shows a partial cross section of the actuator of the third embodiment. It is a principal part overhead view which shows the control mechanism in 4th Embodiment. The variable mechanism in the same embodiment is shown, A is an operation explanatory view showing the closed state of the intake valve by the variable mechanism, B is an operation explanatory view showing the open state of the intake valve by the variable mechanism. It is a longitudinal cross-sectional view which shows the actuator of this embodiment.

  Hereinafter, an embodiment of an actuator according to the present invention will be described in detail with reference to the drawings. In this embodiment, the actuator is applied to a variable valve device on the intake side of a V-type 6-cylinder internal combustion engine, and only one side of the three cylinders is shown in the drawing.

  As shown in FIG. 1, the variable valve operating apparatus is an engine valve that is slidably provided on a cylinder head 1 via a valve guide (not shown) and is urged in a closing direction by valve springs 3 and 3. A pair of intake valves 2, 2, a variable mechanism 4 that variably controls a valve lift amount and an operating angle, which are operating characteristics of the intake valves 2, 2, and a control mechanism 5 that controls the operating position of the variable mechanism 4. And a drive mechanism 6 that rotationally drives the control mechanism 5.

  The variable mechanism 4 includes a hollow drive shaft 13 rotatably supported by a bearing 14 provided on the upper portion of the cylinder head 1, and a drive cam 15 which is an eccentric rotary cam fixed to the drive shaft 13 by press-fitting or the like. And is pivotally supported on the outer peripheral surface of the drive shaft 13 and slidably contacts the upper surfaces of the valve lifters 16, 16 disposed at the upper ends of the intake valves 2, 2 to open the intake valves 2, 2. Transmission that transmits the rotational force of the drive cam 15 to the swing cams 17 and 17 as a swing force linked to the two swing cams 17 and 17 to be operated and the drive cam 15 and the swing cams 17 and 17. And a mechanism.

  The drive shaft 13 is arranged along the longitudinal direction of the engine, and is driven from an engine crankshaft via a driven sprocket (not shown) provided at one end, a timing chain wound around the driven sprocket, and the like. A rotational force is transmitted, and this rotational direction is set in the clockwise direction (arrow direction) in FIG.

  As shown in FIG. 5A, the bearing 14 is provided at the upper end portion of the cylinder head 1 to support the upper portion of the drive shaft 13, and the control shaft described later is provided at the upper end portion of the main bracket 14a. The brackets 14a and 14b are rotatably fastened together by a pair of bolts 14c and 14c.

  The drive cam 15 has a substantially ring shape, and includes an annular cam main body and a cylindrical portion integrally provided on the outer end surface of the cam main body, and a drive shaft insertion hole is formed through the inner shaft. In addition, the axis Y of the cam body is offset from the axis X of the drive shaft 13 in the radial direction by a predetermined amount.

  Both the swing cams 17 have substantially the same raindrop shape and are integrally provided at both ends of the cylindrical camshaft 20, and the camshaft 20 is connected to the drive shaft 13 via the inner peripheral surface. Is supported rotatably. In addition, a pin hole is formed through the tip of the cam nose portion 21 and a cam surface 22 is formed on the lower surface. The cam surface 22 includes a base circle surface on the camshaft 20 side, a ramp surface extending in an arc shape from the base circle surface to the cam nose portion 21 side, and a peak of a maximum lift that is provided on the tip side of the cam nose portion 21 from the ramp surface. The base circle surface, the ramp surface, and the lift surface come into contact with a predetermined position on the upper surface of each valve lifter 16 in accordance with the swing position of the swing cam 17. Yes.

  The transmission mechanism includes a rocker arm 23 disposed above the drive shaft 13, a link arm 24 linking the one end 23 a of the rocker arm 23 and the drive cam 15, the other end 23 b of the rocker arm 23, and the swing cam 17. And a link rod 25 that links the two.

  The rocker arm 23 is rotatably supported by a control cam 33 (to be described later) through a support hole at a cylindrical base portion at the center. Further, the one end portion 23a protruding from the outer end portion of the cylindrical base portion has a pin hole through which the pin 26 is fitted, while the other end portion protruding from the inner end portion of the base portion. 23b has a pin hole into which a pin 27 connected to one end of the link rod 25 is inserted.

  The link arm 24 includes an annular base portion 24a having a relatively large diameter and a projecting end 24b projecting at a predetermined position on the outer peripheral surface of the base portion 24a. The drive cam is located at the center of the base portion 24a. A fitting hole 24c into which 15 cam bodies are rotatably fitted is formed, and a pin hole through which the pin 26 is rotatably inserted is formed in the protruding end 24b.

  The link rod 25 is formed in a substantially square shape having a concave shape on the rocker arm 23 side, and is inserted into each pin hole of the other end portion 23b of the rocker arm 23 and the cam nose portion 21 of the swing cam 17 at both end portions 25a and 25b. Pin insertion holes through which end portions of the pins 27 and 28 are rotatably inserted are formed.

  A snap ring (not shown) that restricts the movement of the link arm 24 and the link rod 25 in the axial direction is provided at one end of each pin 26, 27, 28.

  The control mechanism 5 is rotatably mounted on the same bearing 14 at a position above the drive shaft 13, and is fixed to the outer periphery of the control shaft 32 and is slidably fitted into a support hole of the rocker arm 23. And a control cam 33 serving as a rocking fulcrum of the rocker arm 23.

  The control shaft 32 is disposed in the longitudinal direction of the engine in parallel with the drive shaft 13, and a journal portion at a predetermined position is rotatably supported between the main bracket 14a and the sub bracket 14b of the bearing 14. In addition, the rotation is controlled in the forward or reverse direction by the drive mechanism 6.

  The control cam 33 has a cylindrical shape, and the position of the axis P2 is deviated from the axis P1 of the control shaft 32 by a predetermined amount.

  As shown in FIG. 1, the control shaft 32 is configured such that the maximum rotation position on one side and the maximum rotation position on the other side are regulated by a stopper mechanism 31. The stopper mechanism 31 includes a stopper wall (not shown) protruding from the upper end of the cylinder head 1 and a stopper member 31a fixed integrally to the outer peripheral surface of the control shaft 32. A semicircular concave groove is formed substantially at the center of the upper end portion, and a pair of first and second stopper portions are formed on both upper surfaces of the concave groove.

  On the other hand, the stopper member 31a is integrally joined to the outer peripheral surface of the control shaft 32 with a disc flange-shaped base 31b fitted into the concave groove through a certain clearance, and the base 31b A fan-shaped stopper piece 31c is integrally provided on the upper end surface so as to protrude radially outward. The stopper piece 31c is brought into contact with the first stopper portion in accordance with the rotation of one side of the control shaft 32 at one end side in the circumferential direction so that the control shaft 32 is brought into the rotation control position of the minimum valve lift amount. The first stopper surface 31d to be regulated is provided, and at the other end side, the control shaft 32 is brought into contact with the second stopper portion as the other side of the control shaft 32 rotates, and the control shaft 32 is rotated at the maximum valve lift amount. The second stopper surface 31e is restricted.

  As shown in FIGS. 1 to 4, the drive mechanism 6 includes a housing 35 fixed to the rear end portion of the cylinder head 1, and an electric motor 36 that is a rotational force generating mechanism fixed to one end portion of the housing 35. And a ball screw mechanism 37 that is a conversion mechanism (reduction gear) that is housed in the housing 35 and transmits the rotational force of the electric motor 36 to the control shaft 32, and the opposite side of the housing 35 from the electric motor 36. And a rotation imparting mechanism 39 that imparts a rotational force in a direction to control the intake valves 2 and 2 to the maximum lift amount by the control shaft 32 via the ball screw mechanism 37. .

  The housing 35 is integrally formed of an aluminum alloy material or the like, is disposed along the direction substantially perpendicular to the axial direction of the control shaft 32 inside, and an elongated housing chamber 35a in which a ball screw mechanism 37 is housed. A bulging chamber 35b projecting upward at the upper end of the housing portion 35a and facing the one end 32a of the control shaft 32, and a substantially annular shape separated from the electric motor 36 via a partition wall 35f. Each having an annular chamber 35c. Further, the storage chamber 35a has a circular opening 35d formed at one end in the axial direction, while the annular chamber 35c is closed at the rear end by an end wall 35e.

  The electric motor 36 is constituted by a proportional DC motor, and is fixed in a state where a tip end portion 38a of a substantially cylindrical motor casing 38 seals the opening portion 35d of the storage chamber 35a. Further, as shown in FIG. 1, the electric motor 36 is driven to rotate forward and backward by a control current output from a control unit 40 that detects the operating state of the engine.

  The control unit 40 feeds back detection signals from various sensors such as a crank angle sensor 41, an air flow meter 42, a water temperature sensor 43, and a potentiometer 44 to detect the rotational position of the control shaft 32, and feeds back the current engine. The operating state is detected by calculation or the like, and a control current is output to the electric motor 36.

  The ball screw mechanism 37 includes a ball screw shaft 45 that is an output shaft disposed substantially coaxially with the drive shaft 36 a of the electric motor 36 in the housing chamber 35 a of the housing 35, and a screw on the outer periphery of the ball screw shaft 45. A ball nut 46 which is a moving member to be joined, a linkage arm 47 which is axially connected to one end portion 32a of the control shaft 32 in the bulging chamber 35b, and the linkage arm 47 and the ball nut 46 are linked. The link member 48 is mainly configured.

  In the ball screw shaft 45, a ball circulation groove 49 which is a screw portion having a predetermined width is continuously formed in a spiral shape on the entire outer peripheral surface excluding both end portions 45a and 45b, and one end opening portion of the storage chamber 35a. Both end portions 45a and 45b facing the interior 35d and the annular chamber 35c are rotatably supported by first and second ball bearings 50 and 51, respectively. The outer peripheral surface of the first ball bearing 50 on the electric motor 36 side is press-fitted and fixed inside the one end opening 35d, and the second ball bearing 51 on the distal end side has a small diameter on the other end wall 35e side. It is press-fitted and fixed inside the section.

  Further, the ball screw shaft 45 is serration-coupled so that the distal end portion of the one end portion 45a and the distal end portion of the motor shaft 36a of the electric motor 36 can be axially moved on the same axis by a cylindrical connecting member 52. The rotational force of 36 is transmitted to the ball screw shaft 45, and a slight movement of the ball screw shaft 45 in the axial direction is allowed.

  The ball nut 46 is formed in a substantially cylindrical shape, and a guide groove for continuously holding a plurality of balls 53 is formed in a spiral manner in cooperation with the ball circulation groove 49 on the inner peripheral surface. At the same time, two deflectors for setting the circulating rows of the plurality of balls 53 at two positions in the axial direction of the ball nut 46 are attached. In other words, this deflector guides the balls 53 so as to return to the circulation row again in order to circulate the plurality of balls 53 rolling between the ball circulation grooves 49 and the guide grooves in the same groove. Is.

  The ball nut 46 is applied with a moving force in the axial direction while converting the rotational motion of the ball screw shaft 45 to the ball nut 46 through the balls 53 into linear motion. The ball nut 46 is rotatably provided with a pivot pin 54 connected to one end of the link member 48 at a substantially central position in the axial direction.

  As shown in FIGS. 2 to 4, the linkage arm 47 is formed in a square shape and is integrally fixed to one end portion 32 a of the control shaft 32 by bolts 47 b and 47 b from the axial direction. The other end of the link member 48 is rotatably connected by a pivot pin 55 to a projecting portion 47a projecting from one end thereof.

  The link member 48 is formed in a substantially H shape by press-molding a plate material, and one end of a pair of parallel flat plate-like link portions 48a, 48a is disposed across both sides of the ball nut 46. The other end is rotatably connected to the linkage arm 47 via the balance pin 55, while being rotatably connected to the ball nut 46 via the pivot pin 54.

  Further, as shown in FIG. 2, the ball screw mechanism 37 moves the control shaft 32 at a position where the ball nut 46 moves to the maximum left direction in the drawing as the ball screw shaft 45 rotates in one direction. The lift amount of the intake valves 2 and 2 is controlled to the minimum lift and the operating angle to the minimum operating angle, and as shown in FIG. 3, the ball nut 46 is rotated along with the rotation of the ball screw shaft 45 in the other direction. In the figure, the lift amount of the intake valves 2 and 2 is controlled to the intermediate lift and the operating angle to the intermediate operating angle via the control shaft 32 at a position moved rightward by a predetermined amount. Further, as shown in FIG. 4, the ball nut 46 is moved to the maximum right direction in the drawing along with the rotation of the ball screw shaft 45 in the other direction, and the intake valve 2, via the control shaft 32. The lift amount of 2 is controlled to the maximum lift and the operating angle to the maximum operating angle.

  The ball nut 46 moves to the right by about 10 mm from the maximum leftward position shown in FIG. 2 to the intermediate movement position shown in FIG. 3, and the lead per rotation of the ball screw shaft 45 is about 4 mm. During this time, the ball screw shaft 45 rotates about 2.5 times. Further, since the ball nut 46 moves about 40 mm from the intermediate movement position of the ball nut 46 to the maximum rightward movement position shown in FIG. 4, the ball screw shaft 45 rotates about 10 times. Therefore, at the maximum movement distance in one direction of the ball nut 46, the ball screw shaft 45 rotates about 12 times in total.

  The rotation imparting mechanism 39 has two large and small first and second spur gears 60 and 61 that are rotating members linked to the ball nut 46, and expansion and contraction that urges the first spur gear 60 to rotate in one direction. The torsion spring 62 which is a deformable spring member is mainly configured.

  The first spur gear 60 is formed in a disk shape having a large diameter and a relatively thin wall, and is rotatably supported by a support shaft 63 that is a reversible transmission shaft that is fixedly passed through a fixing hole 60a drilled in the center. In addition, a plurality of tooth portions 60b are formed on the outer periphery.

  The support shaft 63 is supported at both ends in the axial direction by a pair of ball bearings 64 and 65 fixed to holding holes respectively provided in the end wall 35e and the partition wall 35f so as to be rotatable forward and backward.

  The second spur gear 61 has a small diameter and a relatively thin disk shape, and is press-fitted and fixed to the other end portion 45b of the ball screw shaft 45 through a shaft hole 61a drilled in the center. A plurality of tooth portions 61b that mesh with the tooth portions 60b of the first spur gear 60 are formed on the outer periphery.

  The gear ratio (N2 / N1) of the first spur gear 60 and the second spur gear 61 is set to about 1: 3. That is, the second spur gear 61 (ball screw shaft 45) is set to rotate once when the first spur gear 60 rotates, for example, by 1/3.

  The torsion spring 62 is wound around the outer periphery of a cylindrical retainer 66 fixed to the outer periphery of the support shaft 63 with a predetermined gap, and one end 62a is inserted into a small hole formed in the end wall 35e. On the other hand, the other end 62b is inserted and locked in a small hole formed in one side of the first spur gear 60.

  The torsion spring 62 is deformed in the direction in which the ball nut 46 is most contracted in the maximum leftward position (minimum lift) shown in FIG. 2 and increases to the maximum number of 8 turns. That is, the maximum moment is generated at this point.

  When no rotational torque is generated in the electric motor 36, the torsion spring 62 expands and deforms from the aforementioned most contracted diameter deformed state, and the valve spring 3 is moved from both the spur gears 60 and 61 to the ball screw shaft 45. The ball nut 46 is moved to the maximum rightward position (maximum lift) shown in FIG. 4 by overcoming the spring reaction force and applying a rotational force in one direction.

  Hereinafter, the operation of the actuator according to the present embodiment will be described. First, for example, in the low rotation operation region including when the engine is idling, the electric motor 36 is rotated by the control current output from the control unit 40, and the ball screw shaft 45 is rotated by this rotational torque.

  Then, with this rotation, each ball 53 rolls between the ball circulation groove 49 and the guide groove, and the ball nut 46 is linearly moved in the maximum left direction as shown in FIG. . As a result, the control shaft 32 is rotationally driven clockwise by the link member 48 and the linkage arm 47, and the first stopper surface 31d abuts on the first stopper portion to restrict further rotation of the control shaft 32.

  Therefore, in the control cam 33, the shaft center P2 rotates around the shaft center P1 of the control shaft 32 with the same radius as shown in FIGS. 5A and 5B, and the thick portion moves away from the drive shaft 13 upward. . As a result, the other fulcrum 23b of the rocker arm 23 and the pivot point of the link rod 25 move upward with respect to the drive shaft 13, so that each swing cam 17 is connected to the cam nose portion 21 via the link rod 25. The side is forcibly pulled up and the whole is rotated clockwise.

  Therefore, when the drive cam 15 rotates and pushes up the one end portion 23 a of the rocker arm 23 via the link arm 24, the valve lift amount is transmitted to each swing cam 17 and each valve lifter 16 via the link rod 25. However, the lift amount becomes sufficiently small.

  Therefore, in such a low rotation region of the engine, the valve lift amount L1 becomes the smallest as shown in FIG. 7, so that the opening timing of each intake valve 2 is delayed and the valve overlap with the exhaust valve is reduced. For this reason, improvement in fuel consumption and stable rotation of the engine can be obtained. Alternatively, the control shaft 32 may be lifted from the stopper position so that the lift is slightly higher than L1, and the lift amount may be finely controlled.

  In addition, when the engine has shifted to the high engine speed region, the electric motor 36 is rotated in reverse by the control current from the control unit 40, and when this rotational torque is transmitted to the ball screw shaft 45 and rotated, the ball nut is accompanied with this rotation. 46 moves linearly from the position shown in FIG. 2 to the right shown in FIG.

  During the linear movement of the ball nut 46 in the right direction, the spring force of the torsion spring 62 acts as an assist force for the rotational force of the electric motor 36 via the first and second spur gears 60 and 61. The ball nut 46 can quickly move rightward by the rotational force of the motor shaft 36 while overcoming the spring reaction force of the valve spring 3.

  Therefore, as shown in FIG. 4, the control shaft 32 rotates counterclockwise, the second stopper surface 31e abuts on the second stopper portion, and further rotation is restricted.

  As a result, the control cam 33 rotates the axis P2 downward as shown in FIGS. 6A and 6B. For this reason, the rocker arm 23 is now moved toward the drive shaft 13 in its entirety, and the other end 23b presses the cam nose portion 21 of the swing cam 17 downward via the link rod 25, thereby moving the swing cam. 17 is rotated counterclockwise by a predetermined amount.

  Therefore, when the drive cam 15 rotates and pushes up the one end portion 23a of the rocker arm 23 via the link arm 24, the valve lift amount is transmitted to the swing cam 17 and the valve lifter 16 via the link rod 25. The lift amount increases.

  Accordingly, in such a high rotation region, the valve lift amount L3 of each intake valve 2 is maximized as shown in FIG. 7, and the opening timing of each intake valve 2 is advanced and the closing timing is delayed. As a result, the intake charging efficiency is improved and a sufficient output can be secured.

  Next, the operation of the rotation imparting mechanism 39 when the engine is stopped will be described.

  That is, when the engine is stopped by turning off the ignition switch, the rotational torque of the electric motor 36 is not generated. Therefore, the first spur gear 60 is rotated in one direction by the spring force of the torsion spring 62 and the second spur gear is rotated. Rotate 61 in the opposite direction.

  As a result, the ball screw shaft 45 is rotated in one direction, the ball nut 46 is moved to the right in the figure against the spring reaction force of the valve spring, and is stopped at the maximum right direction position shown in FIG. Holds this position stably (default position).

  As a result, the valve lift amount of the intake valves 2 and 2 increases to the maximum lift L3 shown in FIG. 7, and the operating angle is also maximized so that the opening timing is substantially earlier than the piston top dead center (TDC). , The closing timing is sufficiently later than the bottom dead center (BDC).

  Thus, by controlling the intake valves 2 and 2 to the maximum lift, a sufficient lift amount can be secured in the cold state, so that the required intake air amount at the time of cold start with a large engine friction can be satisfied. As a result, good startability can be obtained. In addition, since the maximum lift is controlled, the compression force during the compression stroke of the engine is reduced to some extent so that a so-called compression state is achieved, so that cranking is improved and good startability is obtained.

  Further, as described above, when the engine is stopped, the ball nut 46 can be automatically and reliably moved to the maximum rightward position in the axial direction by the spring force of the torsion spring 62. For example, the electric motor 36 fails. Even when the engine cannot be operated, a good startability of the engine can be ensured, and a fail-safe function works.

  Here, as described in Japanese Patent Application Laid-Open No. 2008-208780, if the urging force is applied to the ball nut, the default position (holding state) is maintained in a rotating state until the engine is completely stopped. Although the lead angle is small when it stops completely, it is difficult to move the ball nut reliably. On the other hand, in the present embodiment, the engine can be reliably moved to the default position even when the engine is stopped.

  When the ball nut 46 moves from the maximum left direction position to the maximum right direction position, the first spur gear 60 rotates twice and the ball screw shaft 45 rotates six times via the second spur gear 61. At this time, the actual number of windings of the torsion spring 62 is deformed by expanding the diameter from 8 windings and decreases by 4 windings to 4 windings. The diameter is expanded at a ratio of 8/4 turns.

  In this way, the torsion spring 62 is reduced by four turns while expanding the diameter until the ball nut 46 is moved to the maximum rightward position. Therefore, the amount of the increased diameter deformation increases as shown in FIG. A gap with the housing 35 is secured. This is because the total number of rotations of the first spur gear 60 is reduced to 1/3 because the gear ratio of the first and second spur gears 60 and 61 is set to 1: 3.

  Therefore, due to the effect of reducing the total number of revolutions, even if the torsion spring 62 is maximally deformed at the maximum lift (maximum rightward position of the ball nut 46), the outer diameter is prevented from expanding and the torsion spring 62 is suppressed. Interference between the outer peripheral edge of 62 and other portions, for example, the housing 35 can be suppressed. As a result, the mountability of the torsion spring 62 is improved.

  That is, in order to move the ball nut 46 from the maximum leftward movement position (intake valve 2, 2 minimum lift) to the maximum rightward movement position (maximum lift), the total number of rotations of the ball screw shaft 45 is 12 rotations. In the case where the torsion spring is directly linked to the ball screw shaft 45 that rotates so much, and the biasing rotational moment is applied, the actual number of windings of the torsion spring at the maximum lift is increased. The number of turns at the minimum lift must be reduced or increased, and the amount of expansion / contraction deformation becomes excessively large. For this reason, there exists a possibility that the mounting | wearing property of a torsion spring may deteriorate. In the former case, the outer diameter becomes excessive, and in the latter case, the axial length becomes long or the spring inner diameter becomes too small.

  However, in the present embodiment, the two first and second spur gears 60 and 61 having different tooth ratios are used, and the torsion spring 62 is linked to the first spur gear 60 having a large diameter. It becomes possible to reduce the rotation speed of the first spur gear 60 to 5 rotations by the gear ratio. Thereby, as described above, the diameter expansion deformation amount and the diameter reduction deformation amount can be made as small as possible. As a result, the mounting property in a narrow space can be improved.

  Further, as described above, since it is not necessary to reduce the total number of rotations of the electric motor while the total number of rotations of the torsion spring 62 is decreased, the motor torque is suppressed, thereby reducing the size and power consumption of the motor. The effect that it can be obtained is also obtained.

  Here, even when the total number of rotations of the torsion spring 62 is not decreased, the degree of freedom of the mounting position of the torsion spring 62 can be improved by appropriately arranging the first and second spur gears 60 and 61. it can. That is, mounting the torsion spring 62 directly to the motor shaft or the ball screw shaft is not easy to mount, but by appropriately arranging the first and second spur gears 60 and 61, a position where the mounting workability is good. It is also possible to arrange a torsion spring 62 on the front.

  The rotation transmission efficiency of each component in the first embodiment, that is, the transmission efficiency associated with the electric motor 36, the control shaft 32, and the ball screw shaft 45, is determined from the electric motor 36. The transmission efficiency from the control shaft 32 to the electric motor 36 is smaller than the transmission efficiency to the shaft 32, the transmission efficiency from the electric motor 36 to the ball screw shaft 45, and the electric power from the ball screw shaft 45 to the electric motor 36. The transmission efficiency to the motor 36 is almost equal.

[Second Embodiment]
FIG. 8 shows a second embodiment of the present invention, and the basic structure is the same as that of the first embodiment. Therefore, the same portions are denoted by the same reference numerals, and detailed description thereof is omitted.

  The disk nut spring retainers 56 and 57 are provided at one end in the axial direction of the ball nut 46 and the inner end of the second ball bearing 51, respectively, and between the spring retainers 56 and 57. A coil spring 58, which is a bias spring for biasing the ball nut 46 in the direction of the electric motor 36 (the control shaft 32 in the direction of the rotation control shaft with the minimum valve lift), is mounted. The coil spring 58 is set to have a relatively small spring load and is balanced with the spring force of the torsion spring 62.

  Therefore, as described above, when the ignition switch is turned off and the engine is stopped, the ball nut 46 tries to move to the maximum left by the spring reaction force of the coil spring 58 and the spring reaction force of the valve spring 3. Since the rotational torque of the electric motor 36 is not generated, the first spur gear 60 rotates in one direction by the spring force of the torsion spring 62 and tries to rotate the second spur gear 61 in the opposite direction.

  As a result, the ball screw shaft 45 rotates in one direction, and the ball nut 46 resists against the spring force of the coil spring 58 and the spring reaction force of the valve spring, and the intermediate position shown in FIG. Move to and hold stably. As a result, the valve lift amount of the intake valves 2 and 2 increases to the intermediate lift L2 shown in FIG. 7, and the operating angle also increases, so that the opening timing is almost the piston top dead center (TDC) position, and the closing timing is Near the bottom dead center (BDC). This intermediate movement position is the lift position of the intake valves 2 and 2 that is optimal for starting the engine, and the lift amount is such that a sufficient intake air amount can be obtained even at the time of cold start with high piston friction. Therefore, the startability of the engine is improved.

  Also in this embodiment, even when the electric motor 36 fails and cannot be operated, a good startability of the engine can be ensured, and the fail-safe function works.

  That is, even when the engine is stopped, the ball screw shaft 45 is urged through the reversible transmission shaft by the spring force of the torsion spring 62, so that it can be moved to the desired ball nut 46 position. .

  In this embodiment, the lift amount of the intake valves 2 and 2 controlled by the variable mechanism 4 is extremely low, for example, about 0.1 mm by the minimum lift control, and the lift amount of the intake valve 2 and 2 is larger than that of the first embodiment by the maximum lift control. This is effective when applied to a variable valve gear set so as to be higher.

  That is, for example, when the maximum lift is set to be relatively high, if the default position when the engine is stopped is set to the maximum lift as in the first embodiment, the friction of the valve train increases and cranking rotation is increased. Although there is a possibility that the rise will be delayed, if the intermediate lift amount is as in the present embodiment, it is possible to reduce the occurrence of excessive friction of the valve system, so that the cranking rotation speed can be increased and good startability can be obtained. It is.

  Conversely, if the minimum lift is set to the default position, the intake air amount necessary for starting cannot be secured at 0.1 mm. Therefore, the intermediate lift is the default. Here, if the minimum lift is about 2 mm, the minimum lift can be a fault. In that case, the torsion spring 62 may be provided so as to be biased toward the minimum lift side.

Further, this embodiment is advantageous in terms of noise and vibration (sound vibration). That is, the ball nut 46 receives a biasing force in the left direction in FIG. 8 by the spring reaction force of the coil spring 58 and the spring reaction force of the valve spring 3, that is, receives a pressing force in the direction of the electric motor 36. On the other hand, a biasing force is applied to the ball nut 46 in the reverse direction by the torsion spring 62. As a result, a gap (backlash gap) between the ball annular groove 49 and the guide groove and the ball 53 can be absorbed to prevent the ball nut 46 from rattling in the axial direction. As a result, backlash noise and vibration can be reduced.
[Third Embodiment]
9A and 9B show a third embodiment, in which a rotation imparting mechanism 39 is disposed outside the housing 35 and the path of the spring biasing force with respect to the output shaft is changed. The structure of the ball screw mechanism 37 and the point having the coil spring 58 are the same as in the second embodiment.

  In the rotation imparting mechanism 39, a motor shaft 36a of the electric motor 36 as an output shaft is formed so as to protrude from the rear end side of the motor casing 38, and the first spur gear 100 with a small diameter fixed to the motor shaft 36a. And a support block 101 bolted to the lower portion of the housing 35, and a pair of support holes 102 and 103 drilled from the lateral direction in the vertical position of the support block 101, and a pair of front and rear ball bearings, respectively. 104a, 104b, 105a, 105b rotatably supported by the first and second support shafts 106, 107, and the large diameter fixed to the tip of the first support shaft 106 and meshed with the first spur gear 100 A second spur gear 108, a small third spur gear 109 fixed to the inside of the second spur gear 108 of the first support shaft 106, and the second support shaft 1 7, a large-diameter fan-shaped fourth spur gear 110 that meshes with the third spur gear 109, and a biasing means 111 that urges the fourth spur gear 110 to rotate in one direction. It is composed of

  The biasing means 111 has a base end portion 112a fixed to the second support shaft 107, a tip end portion 112b formed in a spherical shape, and a sliding hole 101a formed in a lower end portion of the support block 101. And a coil spring 114 that urges the pressure rod 113 in the advancing direction.

  In the pressing rod 113, a flange portion 113 a provided integrally with the distal end portion presses the distal end portion 112 b of the arm 112 in the tangential direction of the fourth spur gear 110 by the spring force of the coil spring 114. 112 is swung within a swing angle range α (conversion angle) shown in the figure. Thus, the fourth spur gear 110 is rotated by the same swing angle range.

  Therefore, according to this embodiment, when the ignition switch is turned off and the engine is stopped, the rotational torque of the electric motor 36 is not generated. Therefore, the ball nut 46 has the spring reaction force of the coil spring 58 and the valve spring. It tries to move to the maximum left direction with the spring reaction force of 3.

  At the same time, on the rotation imparting mechanism 39 side, the pressing rod 113 moves forward through the sliding hole 101a by the spring force of the coil spring 114. As a result, the flange portion 113a at the tip presses the tip portion 112b of the arm 112 and rotates the fourth spur gear 110 in the clockwise direction of FIG. 9B.

  Then, the third spur gear 109 rotates counterclockwise, and the second spur gear 108 rotates counterclockwise in the opposite direction via the first support shaft 106. Therefore, the first spur gear 100 rotates in the same direction (clockwise) as the fourth spur gear 110 to rotate the motor shaft 36a in the same direction. As a result, the ball screw shaft 45 rotates to move the ball nut 46 to the position of about 10 mm in the right direction in the figure against the spring force of the coil spring 58 and the spring reaction force of the valve spring. Stop at the intermediate position shown in Fig. 3 and hold it stably.

  As a result, the valve lift amount of the intake valves 2 and 2 increases to the intermediate lift L2 shown in FIG. 7 as in the second embodiment, and the operating angle increases and the opening timing is almost the piston top dead center (TDC). ) Position, and the closing time is near the bottom dead center (BDC). This intermediate movement position is the lift position of the intake valves 2 and 2 that is optimal for starting the engine, and the lift amount is such that a sufficient intake air amount can be obtained even at the time of cold start with high piston friction. Therefore, the startability of the engine is improved.

  In this embodiment, as shown in FIG. 9A, the mounting position of the urging means 75 can be selected as appropriate by appropriately setting the arrangement of the first and second support shafts 70 and 71.

Furthermore, since the rotation angle (conversion angle) of the second support shaft 107 can also be reduced by the reduction ratio, the extension deformation of the coil spring 114, that is, the stroke of the pressing rod 113 can be shortened. Therefore, the degree of freedom of the layout of the rotation imparting mechanism 39 can be improved.
[Fourth Embodiment]
10 to 12 show a fourth embodiment, in which the specific structure of the variable mechanism 4, the control mechanism 5, and the drive mechanism 6 that is an actuator is changed.

  As shown in FIGS. 11A and 11B, the variable mechanism 4 has an internal hollow shape supported by a large number of partition walls provided at regular intervals on the upper portion of the cylinder head 1 so as to be immovable in the axial direction and the circumferential direction. A rocker shaft 70, a control shaft 71 inserted in the central hole of the rocker shaft 70 so as to be displaceable in the axial direction, and an arm assembly provided on the outer periphery of the rocker shaft 70 and arranged in the same number as the number of cylinders 72 and a drive mechanism 6 (see FIG. 12) for driving the control shaft 71 forward and backward in the axial direction.

  As shown in FIG. 10, the control shaft 71 is disposed in the front-rear direction of the engine, has an outer diameter that is relatively small, and has the drive mechanism 6 at the tip of one end 71a. It is linked.

  The arm assembly 72 is arranged so as to correspond to each cylinder with respect to the rocker shaft 70, and as shown in FIGS. 10 and 11, a drive cam 73a and one end portion provided on the outer periphery of the camshaft 73 on the intake side are provided. An input arm 75, two swing cams 76, and a slider gear 77 are provided between a roller 74 a provided at the center of the swing arm 74 in contact with the stem end of each intake valve 2. .

  The input arm 75 includes an arm portion that protrudes in a direction away from the outer peripheral surface of the rocker shaft 70, and a roller portion 75a that is rotatably provided at the tip of the arm portion and rolls on the outer peripheral surface of the drive cam 73a. It is equipped with.

  Each of the swing cams 76 has a nose portion 76a formed in a substantially triangular shape that protrudes away from the outer peripheral surface of the rocker shaft 70. The roller 74a of the swing arm 74 is pressed against the concave cam surface 76b formed on the lower surface of the nose portion 76a by the spring force of the valve spring 3.

  The input arm 75 and the swing cams 76 integrally swing about the axis of the control shaft 71 (rocker shaft 70), and come into contact with the drive cam 73a when the cam shaft 73 rotates. The input arm 75 swings as the input arm 75 swings via the roller portion 75a. The swinging force of each swing cam 76 is transmitted to each intake valve 2 via the swing arm 74 to open and close.

  The slider gear 77 is supported so as to be rotatable with respect to the outer peripheral surface of the rocker shaft 70 and to be slidable in the axial direction, and a right-handed helical gear 78 is provided at the center in the axial direction. At the same time, two helical gears 79 and 79 are provided on both sides of the helical gear 78, opposite to the helical gear 78.

  On the other hand, on the inner peripheral surfaces of the input arm 75 and the two swing cams 76 that define a space for accommodating the slider gear 77, a right-handed screw and a left-handed screw corresponding to the helical gears 78, 79, 79, respectively. Helical splines are formed and meshed with the helical gears 78, 79, 79, respectively.

  The slider gear 77 is formed with a long hole 80 that extends between the helical gear 78 and one helical gear 79 and extends in the circumferential direction, while the rocker shaft 70 has one of the long holes 80. A long hole (not shown) extending in the axial direction is formed so as to overlap the portion. The control shaft 71 is integrally provided with a locking pin 81 that protrudes through the overlapping portion of the two long holes.

  Therefore, when the control shaft 71 is moved in the axial direction by the drive mechanism 6, the slider gear 77 is pushed by the locking pin 81 and the helical gears 78, 79, 79 are simultaneously moved in the axial direction of the control shaft 71. Since the input arm 75 and the swing cam 76 that are spline engaged with the helical gears 78 and 79 in the axial direction do not move in the axial direction, the shaft of the control shaft 71 is engaged through the meshing of the helical spline. Rotate around the heart.

  At this time, since the input arm 75 and the swing cam 76 have the opposite directions of the formed helical splines, the rotation directions are opposite to each other. As a result, the relative phase difference between the input arm 75 and the swing cam 76 changes, and the lift amount and operating angle of the intake valve 2 are changed.

  As shown in FIG. 12, the driving mechanism 6 is a rotational force generating mechanism accommodated in a housing 82 fixed to the rear end portion of the cylinder head 1 and a motor accommodating chamber 82 a on the inner end side of the housing 82. An electric motor 83, a screw mechanism 84 that is housed in the housing 82 and is a conversion mechanism (reduction gear) that transmits the rotational force of the electric motor 83 to the control shaft 71, and the electric motor 83 of the housing 82 And a rotation applying mechanism 39 that is accommodated in the opposite position and applies a rotational force to the control shaft 71 to control the intake valves 2 and 2 to the large operating angle side via the screw mechanism 84. ing.

  The electric motor 83 is a proportional type DC motor, and is forward / reversely controlled by a control signal from the control unit 40 as in the first embodiment.

  The screw mechanism 84 is fixed to a rotor (not shown) of the electric motor 83, and an output shaft whose front and rear end portions 85 a and 85 b in the axial direction are rotatably supported by ball bearings 86 and 86 provided in the housing 82. A cylindrical member 85 that is a reversible transmission shaft, a drive rod 87 having one end portion 87a inserted and disposed inside the cylindrical member 85, and the other end portion 87b extending to the outside directly connected to the control shaft 71 from the axial direction. , Mainly consists of.

  The cylindrical member 85 has a trapezoidal female thread portion 85c formed on the inner peripheral surface, and a tip end portion 85d of the other end portion 85b is formed in a small stepped diameter. The cylindrical member 85 has an annular groove 85e formed on the outer peripheral surface of the other end portion 85b. On the other hand, the cylindrical member 85 is fitted into the annular groove 85e from the radial direction. A restricting portion 89 that restricts the axial movement of 85 protrudes.

  The drive rod 87 has a trapezoidal male screw portion 87c that meshes with the trapezoidal female screw portion 85c on the outer peripheral surface of the one end portion 87a, and a small diameter tip portion 87d of the one end portion 87a has a substantially U-shaped cross section. The inner support groove 88a is slidably supported in the axial direction. Further, the other end portion 87 b of the drive rod 87 is supported by the guide hole 82 c of the partition wall 82 b integrally provided inside the front end side of the housing 82 so as to be slidable in the axial direction.

  When the cylindrical member 85 is rotated forward or reverse by the electric motor 83, the drive rod 87 is moved in the axial direction through the engagement of the trapezoidal female and male threaded portions 85c and 87c. When the leads of 85c and 87c, that is, the cylindrical member 85 is rotated once, the drive rod 87 is set to move about 2 to 3 mm in the axial direction.

  Further, the drive rod 87 is formed with a long concave groove 87d along the axial direction on the lower surface on the other end 87b side. On the other hand, a restricting piece 90 is provided in the vicinity of the partition wall 82b of the housing 82 so as to restrict the maximum movement stroke in the front-rear direction of the drive rod 87 while the tip edge is in contact with the concave groove 87d. The rotation of the drive rod 87 (control shaft 71) is restricted by the concave groove 87d, and the maximum back and forth movement stroke is set to about 5 mm.

  Therefore, in view of the lead length, when the cylindrical member 85 rotates twice, the valve lift amount of the intake valves 2 and 2 changes from L1 to L3 via the control shaft 71.

  The rotation imparting mechanism 39 is accommodated in an accommodating chamber 82d formed between the motor accommodating chamber 82a and the partition wall 82b, and the basic structure is substantially the same as that of the first embodiment. The same reference numerals are used for simple explanation.

  Two large and small first and second spur gears 60, 61 linked to the cylindrical member 85, and a torsion spring 62, which is a spring member capable of expanding and contracting to urge the first spur gear 60 to rotate in one direction. , Mainly consists of.

  The first spur gear 60 having a large diameter is rotatably supported by a support shaft 63 that passes through and is fixed to a fixing hole 60a drilled in the center, and a plurality of tooth portions 60b are formed on the outer periphery. Yes.

  The support shaft 63 is rotatably supported by a pair of ball bearings 64 and 65 fixed to holding holes respectively provided in the end wall 82b and the partition wall 82e at both ends in the axial direction. A position sensor 91 for detecting the rotational position of the support shaft 63 is provided on the rear end side of the support shaft 63, and rotation angle information detected by the position sensor 91 is output to the controller. The variable mechanism 4 is controlled.

  In the second spur gear 61 having a small diameter, a small-diameter tip end portion 85d of the cylindrical member 85 is inserted and fixed in a shaft hole 61a drilled in the center. A plurality of meshing tooth portions 61b are formed. Note that the second spur gear 61 can be integrally formed with the cylindrical member 85.

  The gear ratio (N2 / N1) between the first spur gear 60 and the second spur gear 61 is set to about 1: 2.5 to 1: 3. Accordingly, the second spur gear 61 (cylindrical member 85) is set to rotate once when the first spur gear 60 rotates, for example, by 1/3. The drive rod 87 (control shaft 71) can control the valve lift amount of the intake valves 2 and 2 from L1 to L3 by a maximum stroke of 5 mm in the axial direction. , At most less than one revolution, ie a rotation angle of less than 360 °.

  The torsion spring 62 is wound around the outer periphery of a cylindrical retainer 66 fixed to the outer periphery of the support shaft 63 with a predetermined gap, and one end 62a is engaged by press-fitting into a small hole formed in the partition wall 82b. On the other hand, the other end 62b is inserted and locked in a small hole formed in one side of the first spur gear 60.

  The torsion spring 62 is deformed in the direction in which the drive rod 87 is contracted to the maximum when the drive rod 87 is in the maximum leftward position (minimum lift) shown in FIG. 12, and increases from the maximum 9 to 10 turns. That is, when the maximum moment is generated at this point and no rotational torque is generated in the electric motor 36, the diameter of the electric motor 36 is increased and the unidirectional rotational force is applied from the spur gears 60 and 61 to the cylindrical member 85. To move the drive rod 87 in the right direction (lift increasing direction).

  Hereinafter, the operation of the present embodiment will be described.

  As described above, when the engine is stopped by turning off the ignition switch, the control shaft 71 (drive rod 87) tends to reach the maximum leftward movement position depending on the individual spring force of each valve spring 3. The rotational torque of the electric motor 83 is not generated, and the first spur gear 60 rotates in one direction by the spring force of the torsion spring 62 and tries to rotate the second spur gear 61 in the opposite direction. For example, the cylindrical member 85 is rotated in one direction, and the drive rod 87 is moved about 1 to 1.5 mm in the right direction in the drawing, stopped at an intermediate position, and stably held. Therefore, the valve lift amount of the intake valves 2 and 2 increases to the intermediate lift L2 shown in FIG. 7, and the operating angle also increases. Therefore, the restartability of the engine is improved as in the first embodiment, and, for example, even when the electric motor 83 fails and cannot be operated, the engine can be started with good fail-safe function. Work.

  In particular, even in the case of a nonreciprocal transmission mechanism that cannot rotate the cylindrical member 85 by moving the control shaft in the axial direction as in the present embodiment, it is possible to shift to the desired intermediate lift L2. In addition, even when the engine is stopped, it is possible to shift to the intermediate lift L2.

  Further, when the drive rod 87 moves to the maximum right from the intermediate movement position, the outer diameter of the torsion spring 62 does not become excessively large although the amount of diameter expansion deformation increases. This is because the number of rotations of the torsion spring 62 is reduced because the gear ratio of the first and second spur gears 60 and 61 is set from 1: 2.5 to 1: 3.

  Therefore, even when the torsion spring 62 is deformed to the maximum at the time of the maximum lift (maximum rightward position of the drive rod 87), the outer diameter of the torsion spring 62 can be suppressed from becoming excessively large. And other parts can be suppressed, and contact with the outer peripheral surface of the shaft on the inner diameter side can be avoided. As a result, the mountability of the torsion spring 62 is improved.

  In addition, in this embodiment, the movement stroke between the maximum left position (minimum lift position) and the maximum right position (maximum lift position) of the drive rod 87 is about 5 mm. Since the rotation angle of the spur gear 60 (support shaft 63) is less than 360 °, the detection accuracy of the valve lift amount of the intake valves 2 and 2 by the position sensor 91 is increased.

  That is, if the first spur gear 60 exceeds a maximum of one rotation (360 °), the rotary position sensor 91 cannot be used, and the axial stroke position of the control shaft 71 must be detected. Therefore, the detection accuracy is likely to be lowered. However, in this embodiment, the rotational position of less than 360 ° has only to be detected by the general position sensor 91, so that the detection accuracy is increased.

  The present invention is not limited to the configuration of each of the above-described embodiments. In each of the above-described embodiments, the ball screw mechanism is shown as the speed reducer. May be. For example, a configuration described in Japanese Patent Application Laid-Open No. 2008-286188 may be configured to change the phase of the worm wheel by rotationally driving a worm with a motor shaft. In particular, in the case of a nonreciprocal mechanism such as a worm wheel, it is difficult to move the motor shaft side from the control shaft side to convert it to an intermediate working angle when the engine is stopped. Can be converted.

  The drive mechanism 6 may be driven by hydraulic pressure in addition to the electric motor, and the actuator of the present application may be applied to the exhaust valve side or both valve sides in addition to the intake valve side. is there.

  Furthermore, the spring member can be applied to various spring members such as a spring spring in addition to the torsion spring and the coil spring.

The technical ideas of the invention other than the claims ascertained from the embodiment will be described below.
[Claim a] In the actuator according to claim 1,
An actuator characterized in that rotation of the output shaft is decelerated and rotational force is transmitted to the reversible transmission shaft.

According to the present invention, even when the output shaft is rotated many times by the electric motor, the rotation is decelerated and transmitted to the reversible transmission shaft. Since the means can be reduced in size, mountability to the apparatus is improved.
[Claim b] In the actuator according to claim a,
The actuator, wherein the reversible transmission shaft is rotated so that the rotation of the output shaft is decelerated so that the reversible transmission shaft rotates within 360 degrees.

According to this invention, since the rotation of the reversible transmission shaft is 1 rotation or less, the mountability to a device such as a torsion spring as the biasing means, that is, the assembly performance is improved.
[Claim c] In the actuator according to claim b,
An actuator, wherein the moving position of the control shaft is detected by detecting the rotational position of the reversible transmission shaft.

According to the present invention, since the full conversion angle of the reversible transmission shaft is within 360 °, the rotational position detecting means can easily detect the movement (rotation) position of the control shaft, that is, the lift amount of the engine valve. In particular, in the case of an axially movable control shaft, the detection method can be further simplified because it can be detected not as an axial direction but as a rotational phase position.
[Claim d] In the actuator according to claim 1,
The actuator, wherein the urging means urges the reversible transmission shaft so that the control shaft reaches a maximum lift amount of the engine valve.

According to the present invention, since a sufficient lift amount can be ensured in the cold state, it is possible to satisfy the required intake air amount at the time of cold start with large engine friction, and as a result, good startability can be obtained.
(Claim e) In the actuator according to claim 1,
Power is transmitted from the output shaft to the control shaft, but power is transmitted from the output shaft to the control shaft by a non-reversible reduction mechanism that does not transmit power from the control shaft to the output shaft. An actuator characterized by.

Even if the speed reduction mechanism is irreversible, it is possible to sufficiently energize in the direction of increasing the lift amount.
[Claim f] In the actuator according to claim 1,
A screw shaft to which a rotational force is transmitted from the output shaft;
A moving member that moves in the axial direction by rotation of the screw shaft,
An actuator configured to move the control shaft by movement of the moving member.
[Claim g] In the actuator according to claim f,
The reversible transmission shaft is an shaft transmitted from the screw shaft through a gear.
(Claim h) In the actuator according to claim g,
The gear provided on the screw shaft is a small-diameter gear, and the gear provided on the reversible transmission shaft is a large-diameter gear.
(Claim i) In the actuator according to claim h,
The urging means is a torsion spring, and one end of the torsion spring is fixed to the large-diameter gear, and the other end of the torsion spring is fixed to a fixed portion.
[Claim j] In the actuator according to claim f,
The actuator according to claim 1, wherein the moving member is biased in a direction opposite to a direction biased by the biasing means.
(Claim k) In the actuator according to claim 1,
An actuator characterized in that the lift amount of the engine valve is varied by rotating the control shaft.
[Claim 1] In the actuator according to claim 1,
An actuator characterized in that the lift amount of the engine valve is varied by moving the control shaft in the axial direction.

2 ... Intake valve (engine valve)
DESCRIPTION OF SYMBOLS 4 ... Variable mechanism 5 ... Control mechanism 6 ... Drive mechanism 32 * 71 ... Control shaft 33 ... Control cam 36 * 83 ... Electric motor 36a ... Motor shaft (output shaft)
37 ... Ball screw mechanism (reduction gear)
39 ... Rotation imparting mechanism 45 ... Ball screw shaft (reversible transmission shaft)
46 ... Ball nut (moving member)
60 ... first spur gear 61 ... second spur gear 62 ... torsion spring (biasing means)
63 ... Support shaft (reversible transmission shaft)
84 ... Screw mechanism (reduction gear)
85 ... Cylindrical member (output shaft / reversible transmission shaft)
87 ... Screw shaft 111 ... Biasing means

Claims (3)

An actuator that variably controls the position of the control shaft in a variable valve gear that varies the operating characteristics of the engine valve by moving the control shaft,
An electric motor that rotationally controls the output shaft according to the state of the internal combustion engine and applies an operating force to the control shaft;
A rotational force is transmitted from the output shaft, and a reversible transmission shaft having reversible characteristics;
Biasing means for imparting biasing force to the reversible transmission shaft;
An actuator comprising: An actuator of a variable valve operating device that varies the operating characteristics of an engine valve by moving a control shaft by the power of an electric motor according to the state of the internal combustion engine ,
Having a reversible transmission shaft that is rotationally driven by the electric motor and in which an urging force is applied in one direction;
The transmission efficiency from the control shaft to the electric motor is smaller than the transmission efficiency from the electric motor to the control shaft,
An actuator characterized in that a transmission efficiency from the electric motor to the reversible transmission shaft is substantially equal to a transmission efficiency from the reversible transmission shaft to the electric motor. A drive shaft in which a rotational force is transmitted from the crankshaft and a drive cam is provided; and
A transmission mechanism for transmitting the power by converting the rotational force of the drive shaft into a swinging force;
A swing cam for operating the engine valve by transmitting the swing force of the transmission mechanism and performing swing motion;
A control shaft for changing the lift amount of the engine valve by changing the transmission path of the transmission mechanism to change the swing state of the swing cam by moving according to the state of the internal combustion engine;
A variable mechanism with
An electric motor that controls the rotation of the output shaft in accordance with the state of the internal combustion engine;
A reversible transmission shaft to which the rotational force of the output shaft is transmitted and having reversible characteristics;
Biasing means for imparting biasing force to the reversible transmission shaft;
An actuator with
A variable valve operating apparatus for an internal combustion engine, comprising:

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