Valve control apparatus for internal combustion engine

A valve control apparatus includes first and second engine valves; a first drive cam configured to rotate integrally with the drive shaft; a second drive cam provided on the drive shaft and configured to rotate integrally with the drive shaft; a swing cam configured to swing; a transmission mechanism configured to convert a rotation of the first drive cam into a swinging force and to transmit the swinging force to the swing cam; a first swing arm configured to open the first engine valve by a swing of the swing cam; a second swing arm configured to open the second engine valve by a rotation of the second drive cam; a control mechanism configured to vary a swing amount of the swing cam by varying an attitude of the transmission mechanism; and a connection changeover mechanism configured to connect and disconnect the first swing arm with/from the second swing arm.

BACKGROUND OF THE INVENTION

The present invention relates to a valve control apparatus for an internal combustion engine, which is able to vary a characteristic such as a lift amount of intake valve and/or exhaust valve in accordance with an operating state of the engine.

Japanese Patent Application Publication No. 2009-103040 discloses a previously-proposed valve control apparatus in the field. This valve control apparatus includes a holder which varies its swing position by a control cam, and a sub-cam which is driven by an intake cam and which swings about a support shaft fixed to the holder. The sub-cam includes a drive cam surface and a rest cam surface. The drive cam surface drives a first intake valve through a first rocker arm. The rest cam surface drives a second intake valve through a second rocker arm. Moreover, the valve control apparatus further includes a connection changeover mechanism which connects the first rocker arm with the second rocker arm or disconnects the first rocker arm from the second rocker arm.

In a high-load region of engine, the connection changeover mechanism connects the first rocker arm with the second rocker arm so that both of the first and second intake valves are driven (opened and closed) by the drive cam surface which produces a large lift. Thereby, an intake-air charging efficiency is enhanced to increase an output torque of the engine.

On the other hand, in a low-load region of the engine, the connection changeover mechanism disconnects the first rocker arm from the second rocker arm. Thereby, the first intake valve is driven by the drive cam surface, and the second intake valve is made substantially in a closed state (minute-lift state) by the rest cam surface which produces a small lift. Because of this lift difference between the first and second intake valves, an intake-air swirl effect is produced in a cylinder, so that a combustion of the engine is improved. Hence, a fuel economy is improved.

SUMMARY OF THE INVENTION

However, in the above-mentioned previous valve control apparatus, lift characteristics of the first and second intake valves vary in conjunction with each other in a case that the swing position of the holder is varied by controlling a phase of the control cam under the unconnected state between both the rocker arms.

That is, because the drive cam surface and the rest cam surface which drive the respective first and second rocker arms are formed together in the sub-cam, both the cam surfaces operate with the same swing-operating characteristic.

As a result, as shown inFIG. 9of the above valve control apparatus, when a working angle (corresponding to a valve-open period) of the first intake valve which produces the large lift is varied, a working angle of the second intake valve which produces the small lift is varied subordinately in conjunction with the variation of the working angle of the first intake valve. Thereby, various inconveniences are caused. For example, when the working angle of the second intake valve has become relatively small, a function to enable fuel and contamination stored at an upper surface of an umbrella portion of the second intake valve during a valve-closed period to be removed during the valve-open period is weakened. Hence, there is a risk that a time-dependent change of combustion is caused. On the other hand, when the working angle of the second intake valve has become relatively large, there is a risk that the swirl function is weakened to worsen the combustion. Moreover, there is a risk that a friction of valve system is increased to worsen a fuel economy.

It is therefore an object of the present invention to provide a valve control apparatus devised to solve or ease the above-mentioned problem.

According to one aspect of the present invention, there is provided a valve control apparatus for an internal combustion engine, comprising: a first engine valve biased in a closing direction of the first valve by a biasing force of a valve spring; a second engine valve biased in a closing direction of the second valve by a biasing force of a valve spring; a first drive cam provided on a drive shaft and configured to rotate integrally with the drive shaft, the drive shaft being configured to rotate in synchronization with a crankshaft; a second drive cam provided on the drive shaft and configured to rotate integrally with the drive shaft; a swing cam configured to swing; a transmission mechanism configured to convert a rotational motion of the first drive cam into a swinging force and to transmit the swinging force to the swing cam; a first swing arm configured to open the first engine valve by being pressed by a swing of the swing cam; a second swing arm configured to open the second engine valve by being pressed by a rotation of the second drive cam; a control mechanism configured to vary a swing amount of the swing cam by varying an attitude of the transmission mechanism; and a connection changeover mechanism configured to connect and disconnect the first swing arm with/from the second swing arm.

According to another aspect of the present invention, there is provided a valve control apparatus for an internal combustion engine, comprising: a first drive cam configured to be rotated drivingly by a rotational force of a crankshaft; a second drive cam configured to be rotated drivingly by the rotational force of the crankshaft; a first engine valve biased in a closing direction of the first valve by a valve spring; a second engine valve biased in a closing direction of the second valve by a valve spring; a transmission mechanism configured to convert a rotational motion of the first drive cam into a swinging motion and to transmit the swinging motion to a swing cam; a control mechanism configured to vary a swing amount of the swing cam by varying an attitude of the transmission mechanism; a first follower configured to open and close the first engine valve by a contact with the swing cam; a second follower configured to open and close the second engine valve by a contact with the second drive cam; and a changeover mechanism configured to form an interlock between opening amount and open-close timing of the first follower and opening amount and open-close timing of the second follower, and configured to release the interlock.

According to still another aspect of the present invention, there is provided a valve control apparatus for an internal combustion engine, comprising: a pair of engine valves including a first engine valve and a second engine valve; a first follower configured to drivingly open and close the first engine valve; a second follower configured to open and close the second engine valve; a first drive cam configured to rotate in synchronization with a crankshaft; a swing cam configured to drivingly press the first follower; a transmission mechanism configured to convert and transmit a rotational motion of the first drive cam to a swinging motion of the swing cam; a control mechanism configured to vary a transfer characteristic of the transmission mechanism by varying an attitude of the transmission mechanism; a second drive cam configured to rotate in synchronization with the crankshaft and to drive the second follower; and a changeover mechanism configured to switch between an interlocked state of the first follower and the second follower and a non-interlocked state of the first follower and the second follower.

DETAILED DESCRIPTION OF THE INVENTION

Hereinafter, embodiments of valve control apparatus for internal combustion engine according to the present invention will be described referring to the drawings. In each embodiment, the valve control apparatus is applied to an intake side and/or an exhaust side of multi-cylinder internal combustion engine.

As shown inFIGS. 1 and 2, a valve control apparatus in a first embodiment according to the present invention includes first and second intake valves3aand3b, a drive shaft4, a swing mechanism6, a single swing cam7, a first drive cam5, a transmission mechanism8, and a control mechanism9. Each of the first and second intake valves3aand3bis provided slidably in a cylinder head1through a valve guide (not shown), and opens and closes an intake port. Each cylinder of the plurality of cylinders is equipped with the first and second intake valves3aand3b, i.e., two engine valves. The drive shaft4is disposed in a front-rear direction of the engine, and is formed in an internally hollow shape. The swing mechanism6is provided on upper end portions of the respective intake valves3aand3b. The single swing cam7operates opening/closing movements of, in principle, the first intake valve3athrough the swing mechanism6. The after-explained first drive cam5is provided on an outer circumference of the drive shaft4. The transmission mechanism8links or coordinates the first drive cam5with the swing cam7. The transmission mechanism8converts a rotational force of the first drive cam5to a swinging motion, and transmits this swinging motion to the swing cam7as a swinging force. Thus, the control mechanism9controls the first intake valve3aso as to continuously vary a valve lift-amount characteristic of the first intake valve3aand a valve working angle (valve-open-period angle range) of the intake valve3ain accordance with an operating state of the engine, by varying an attitude (position) of the transmission mechanism8and thereby varying a swing range of the swing cam7.

In this embodiment, the valve working angle means a time interval for which each intake valve3a,3bis open. Moreover, the swing cam7cooperates with the transmission mechanism8and the control mechanism9to define a variable mechanism. This variable mechanism is provided to every cylinder. That is, each cylinder has one variable mechanism which is constituted by the swing cam7, the transmission mechanism8and the control mechanism9.

The first intake valve3ais biased (urged) by a valve spring10ain a direction that closes (blocks) an open end of the intake port. The valve spring10ais resiliently attached between a bottom portion of an approximately-cylindrically-shaped bore formed in an upper end portion of the cylinder head1and a spring retainer provided to an upper end portion of valve stem. In the same manner, the second intake valve3bis biased by a valve spring10bin a direction that closes or blocks an open end of the intake port. The valve spring10bis resiliently attached between a bottom portion of an approximately-cylindrically-shaped bore formed in the upper end portion of cylinder head1and a spring retainer provided to an upper end portion of valve stem.

Predetermined axial portions and both end portions of the drive shaft4are rotatably supported by first and second bearing portions11aand11band bearing portions11c. The first and second bearing portions11aand11bare provided in an upper portion of the cylinder head1and are arranged on both lateral portions of the variable mechanism. Each cylinder includes one pair of first and second bearing portions11aand11b. The bearing portions11care provided on the both end portions of the drive shaft4. The drive shaft4is formed with an oil passage provided axially inside the drive shaft4. Lubricating oil passed through the oil passage is supplied to the respective bearing portions11ato11cand the like. The first drive cam5is fixed to a predetermined axial portion of the outer circumference of the drive shaft4. Moreover, a second drive cam13is provided at a location axially separated from (axially away from) the first drive cam5. Every cylinder includes one first drive cam5and one second drive cam13.

Moreover, a timing chain (not shown) is provided on one end portion of the drive shaft4, and thereby, rotational force is transmitted from a crankshaft of the engine through the timing chain to the drive shaft4. Thus, the drive shaft4is able to rotate in a clockwise direction (arrow direction) ofFIG. 1.

The first drive cam5includes a cam main body5aand a boss portion5b. The cam main body5ais formed approximately in a disc shape. As shown inFIG. 2, the boss portion5bis formed in a tubular shape, and is provided integrally with an (axially) outside portion of the cam main body5a. The first drive cam5is fixed to the drive shaft4by a fixing pin12. The fixing pin12passes through a pin hole which was drilled in the boss portion5bin a radial direction. Moreover, the first drive cam5is disposed on one end side (i.e., on one lateral side) of the swing cam7relative to an axial direction of the drive shaft4. The boss portion5bis located on an opposite side of the cam main body5afrom the swing cam7. An outer circumferential surface of the cam main body5ais formed in a cam profile of eccentric circle. That is, a shaft center X (i.e., a center of the outer circumferential surface) of the cam main body5ais offset (deviated) from a shaft center Y of the drive shaft4in the radial direction by a predetermined amount.

As shown inFIGS. 1 and 4C, the second drive cam13is formed by cutting an outer circumferential surface of the drive shaft4along a circumferential direction of the drive shaft4. An outer circumferential surface13aof the second drive cam13is formed in a circular (annular) shape having a small diameter in cross section taken by a plane perpendicular to the axial direction such that the outer circumferential surface13ais constituted as a so-called oval cam (egg-shaped cam). The entire outer diameter of the second drive cam13is smaller than an outer diameter of the drive shaft4. The outer circumferential surface13aof the second drive cam13includes a base circular portion and a cam nose portion13bas shown inFIG. 4C. When the second drive cam13rotates in synchronization with the derive shaft4, the base circular portion and the cam nose portion13bof the outer circumferential surface13aopen and close the second intake valve3bthrough an after-mentioned second swing arm31of the swing mechanism6.

As shown inFIG. 1, the swing mechanism6is constituted by two of a first swing arm30functioning as a first follower and the second swing arm31functioning as a second follower. The second swing arm31is provided adjacent to a lateral portion of the first swing arm30relative to the axial direction. The both swing arms30and31are provided independently from each other (i.e., provided as components that can move independently from each other). The first swing arm30includes a base end portion30aand a tip portion30b, and the second swing arm31includes a base end portion31aand a tip portion31b. The base end portions30aand31aare swingably supported by one rocker shaft32. The tip portions30band31bprotrude in the same direction respectively from the base end portions30aand31a. A lower surface of the tip portion30bis formed with a circular concave portion. Similarly, a lower surface of the tip portion31bis formed with a circular concave portion. The tip portion30bis in contact with the upper surface of a stem end of first intake valve3athrough a disc-shaped shim33afitted into the concave portion of tip portion30b. Similarly, the tip portion31bis in contact with the upper surface of a stem end of second intake valve3bthrough a disc-shaped shim33bfitted into the concave portion of tip portion31b.

The first swing arm30is provided at a location identical with a location of the swing cam7relative to a width direction of the engine (right-left direction ofFIG. 4A). A roller34is provided to an approximately center portion of width range of the first swing arm30relative to the axial direction of rocker shaft32. The roller34rotatably abuts on an after-mentioned cam surface of the swing cam7. An approximately center portion of this roller34relative to a width direction of roller34accords with the location of an axis (stem center) Z of the valve stem of first intake valve3a. The roller34is rotatably received by a concave groove of the first swing arm30through a roller shaft34a. This concave groove is formed at an approximately center portion of the first swing arm30. An upper end portion of the roller34is constantly exposed to the side of swing cam7.

The second swing arm31is provided to be offset from (away from) the swing cam7in the axial direction. Hence, the swinging force of swing cam7is not directly transmitted to the second swing arm31. A spherical lower surface of the shim33bfitted in the tip portion31bis in contact with the upper surface of stem end of second intake valve3b. When an after-mentioned connection changeover mechanism36connects (interlocks) the second swing arm31with the first swing arm30, the second swing arm31largely opens the second intake valve3bby pressing against a spring force of the valve spring10b.

The second swing arm31includes a slip convex portion35at an approximately center portion of the second swing arm31relative to a width direction of the second swing arm31. That is, the slip convex portion35is formed integrally with the second swing arm31to protrude from an upper surface of the second swing arm31. The slip convex portion35is formed in an approximately rectangular shape as viewed from the axial direction of the rocker shaft32. The slip convex portion35has a slip surface35aas an upper surface of the slip convex portion35. When the second swing arm31is swinging, the slip surface35aof the slip convex portion35is elastically in contact with the outer circumferential surface13aof the second drive cam13in the radial direction of the second drive cam13by the biasing force of the valve spring10b.

The respective lower surfaces of shims33aand33bwhich are in contact with the first and second intake valves3aand3bare formed in an approximately spherical shape. Thereby, when each swing arm30,31swings, the shim33a,33bcan press a portion near the center (line Z ofFIGS. 1 and 2) of stem end of the intake valve3a,3b.

Moreover, a thickness of the shim33ais appropriately selected by selecting from multiple shims having different thickness values, so that a space between the stem end of first intake valve3aand the shim33ais adjusted to become a slight clearance near zero especially when the first intake valve3ais in a non-lifted state (closed state). Similarly, the shim33bis appropriately selected among multiple shims having different thickness values, so that the a space between the stem end of second intake valve3band the shim33bis adjusted to become a slight clearance near zero when the second intake valve3bis in the non-lifted state (the closed state) under a state where the both swing arms30and31have been connected (interlocked) with each other by the after-mentioned connection changeover mechanism36.

As shown inFIG. 2, the connection changeover mechanism36includes a first retaining hole37a, a second retaining hole37b, a connecting pin38, a coil spring39, a pressure-receiving chamber40, and a hydraulic circuit41. The second swing arm31is formed with the first retaining hole37awhich functions as a connection hole of the second swing arm31. The first swing arm30is formed with the second retaining hole37bwhich functions as a connection hole of the first swing arm30. The first retaining hole37aand the second retaining hole37bare formed continuously inside the both base end portions30aand31aof swing arms30and31in the axial direction. The connecting pin (connecting member)38is provided for the interlock between the first and second swing arms30and31, and is retained in the first retaining hole37a. A front-end portion38aof the connecting pin38can slide into the second retaining hole37bso as to engage the first swing arm30with the second swing arm31. The coil spring39is elastically retained in the second retaining hole37b, i.e., is a biasing member for biasing the connecting pin38toward the first retaining hole37a. The pressure-receiving chamber40is formed on a rear-end side of the first retaining hole37a. The pressure-receiving chamber40can apply oil pressure to the connecting pin38to appropriately move the connecting pin38toward the second retaining hole37bagainst the biasing force of coil spring39. The hydraulic circuit41supplies/discharges oil pressure to/from the pressure-receiving chamber40.

The hydraulic circuit41includes a hydraulic-pressure supply/discharge passage43, an oil pump44, an electromagnetic changeover valve48, and an electronic controller (ECU)49. As shown inFIG. 2, the hydraulic-pressure supply/discharge passage43supplies and discharges working oil pressure to/from the pressure-receiving chamber40through an oil hole42aand an oil passage42. The oil passage42is formed axially inside the rocker shaft32. The oil pump44pumps working oil stored in an oil pan45, through a supply passage46to the hydraulic-pressure supply/discharge passage43. The electromagnetic changeover valve48switches between the supply passage46and a drain passage47in order to communicate one of the supply passage46and the drain passage47with the hydraulic-pressure supply/discharge passage43. The electronic controller49controls the switching operation of electromagnetic changeover valve48.

The electronic controller49receives information signals derived from various kinds of sensors such as a crank angle sensor, an air flow meter and an engine water-temperature sensor (not shown). Thus, the electronic controller49detects a current operating state of the engine, and thereby, outputs control signals to the electromagnetic changeover valve48.

As shown inFIGS. 1 and 2, the swing cam7is formed approximately in a raindrop shape. The swing cam7is formed integrally with a cam shaft7aprovided on a side of base end portion of swing cam7. The cam shaft7ais formed in a short circular-tube shape, and is fitted over the outer circumferential surface of drive shaft4by insertion. The swing cam7is supported to be able to swing about the shaft center Y of drive shaft4via the cam shaft7a. That is, the shaft center Y serves as a swing axis of the swing cam7. (FIG. 4A)

The swing cam7includes a cam nose portion7bin a tip side of the swing cam7. As shown inFIG. 4A, a lower surface of the swing cam7includes a cam surface7dformed between the base end portion of the swing cam7and the cam nose portion7b. This cam surface7dincludes a base circular surface, a ramp surface and a lift surface. The base circular surface is located at a side of the base end portion. The ramp surface extends in a circular-arc shape (in cross section) from the base circular surface toward the cam nose portion7b. The lift surface extends from the ramp surface to a maximum-lift top surface of the cam surface7d. This maximum-lift top surface is located in a tip side of the cam nose portion7b. The cam surface7dis in contact with the outer circumferential surface of the roller34of the first swing arm30. The swing cam7varies the lift amount of the intake valve3a,3b, by varying a contact point between the cam surface7dand the roller34in accordance with a swing position of the swing cam7.

A swinging direction of swing cam7when opening the first intake valve3a(i.e., when the contact point between the cam surface7dand the roller34moves toward the lift surface) is identical with a rotational direction of the drive shaft4(arrow direction inFIG. 1). Accordingly, a drag torque is applied to the swing cam7in the direction that lifts the first intake valve3a, because of a friction coefficient between the drive shaft4and the swing cam7. Therefore, a drive efficiency of the swing cam7is improved.

Moreover, the swing cam7includes a connecting portion7clocated on an opposite side of the cam shaft7afrom the cam nose portion7b. That is, the cam shaft7ais located between the cam nose portion7band the connecting portion7c, and this connecting portion7cis formed integrally with the swing cam7to protrude from the swing cam7. The connecting portion7cis formed with a pin hole passing through both lateral surfaces of the connecting portion7c, i.e., passing through the swing cam7in the axial direction of drive shaft4. A connecting pin18for connecting the swing cam7with an after-mentioned another end portion17bof link rod17is inserted into the pin hole.

As shown inFIGS. 1 to 4C, the transmission mechanism8includes a rocker arm15, a link arm16and the link rod17. The rocker arm15is disposed (to extend) along the width direction of engine above the drive shaft4. The link arm16links the rocker arm15with the drive cam5. The link rod17links the rocker arm15with the connecting portion7cof swing cam7. That is, the transmission mechanism8is constructed as a mechanical multi-joint link mechanism including the rocker arm15, the link arm16and the link rod17.

As shown inFIGS. 3A and 3B, the rocker arm15includes a tubular base portion15a, a first arm portion15band a second arm portion15c. The tubular base portion15ais located in one end side of the rocker arm15, and is swingably supported by an after-mentioned control eccentric shaft29. The first and second arm portions15band15care located in another end side of the rocker arm15, and are provided to protrude approximately parallel to each other from an outer surface of the tubular base portion15atoward an inside of the engine, in a biforked manner.

The tubular base portion15ais formed with a support hole15dpassing through the tubular base portion15a. The tubular base portion15ais supported by causing the support hole15dto be fitted over an after-mentioned outer circumference of the control eccentric shaft29through a minute clearance therebetween.

The first arm portion15bis formed integrally with a shaft portion15ethat protrudes from an outside surface of tip portion of the first arm portion15b. The shaft portion15eis linked rotatably with an after-mentioned protruding end16bof the link arm16.

On the other hand, the second arm portion15cincludes a block portion15fat a tip portion of second arm portion15c. A lift adjusting mechanism21is provided to the block portion15f. One end portion17aof the link rod17is linked rotatably with an after-mentioned pivotally-supporting pin19of the lift adjusting mechanism21.

Moreover, the block portion15fis formed with an elongate hole (slot hole)15hpassing through the block portion15fin a lateral direction of the block portion15f. That is, the elongate hole15his formed to pass from one side of block portion15fto another side of block portion15fin the axial direction of drive shaft4. The pivotally-supporting pin19laterally inserted in the elongate hole15his capable of moving within the elongate hole15hin an upper-lower direction, i.e., moving along the elongate shape of hole15h, for adjustment.

The first arm portion15band the second arm portion15care provided to have angles different from each other in a swinging direction of the rocker arm15. That is, there is some angle between an imaginary linkage center line of the first arm portion15band an imaginary linkage center line of the second arm portion15c. Also, the first arm portion15band the second arm portion15care positioned to deviate from each other in the upper-lower direction. The tip portion of first arm portion15bis more inclined toward the lower direction by a slight inclination angle than the tip portion of the second arm portion15c.

As shown inFIGS. 1,2and4B, the link arm16includes an annular portion (circular tube portion)16aand the protruding end16b. The annular portion16ahas a relatively large diameter. The protruding end16bis provided to protrude from a predetermined portion of outer circumferential surface of the annular portion16a. A fitting hole16cis formed at a center portion of the annular portion16a. The fitting hole16cis rotatably fitted over an outer circumferential surface of the cam main body5aof the drive cam5so that the drive cam5rotatably supports the link arm16.

The link rod17includes both rod portions located away from each other in the axial direction of drive shaft4. These two rod portions are integrally formed by press molding. Hence, the link rod17is shaped like a U-shape in cross section. In order to attain a compactification inside the link rod17, the link rod17is formed by being bent in an approximately circular-arc shape. The one end portion17a(of each rod portion) of link rod17is connected with the second arm portion15cthrough the pivotally-supporting pin19inserted into a pin hole of the one end portion17a. The another end portion17bof link rod17is connected rotatably with the connecting portion7cof the swing cam7through the connecting pin18inserted into a pin hole of the another end portion17b. Moreover, since only one link rod17is provided to each cylinder of the engine, a structure of the valve control apparatus can be simplified while lightening a weight of the apparatus.

The swing cam7swings in the lifting direction when the link rod17raises (pulls up) the connecting portion7c. Since the cam nose portion7bthat receives an input from the roller34is located on the opposite side of a swinging center of swing cam7from the connecting portion7c, a generation of fall (inclination) of swing cam7can be suppressed.

As shown inFIGS. 1 and 2, the lift adjusting mechanism21includes the pivotally-supporting pin19, an adjusting bolt22, and a lock bolt23. The pivotally-supporting pin19is provided in the elongate hole15hof block portion15fof second arm portion15cof rocker arm15. The adjusting bolt22is screwed into an adjusting female threaded hole from its lower side. This adjusting female threaded hole is drilled in a lower portion of the block portion15ftoward the elongate hole. Moreover, a fixing female threaded hole is drilled in an upper portion of the block portion15ftoward the elongate hole. The lock bolt23is screwed into the fixing female threaded hole from its upper side.

After an assembling of the respective structural elements, a fine adjustment for the lift amount of each intake valve3a,3bis carried out by adjusting an up-down position of the pivotally-supporting pin19within the elongate hole15h(a position set along elongate shape of the elongate hole15h) by use of the adjusting bolt22. After this fine adjustment, the position of pivotally-supporting pin19is fixed (fastened) by tightening the lock bolt23.

The control mechanism9includes a control shaft24and an electric actuator (not shown). The control shaft24is disposed parallel to the drive shaft4, in a region above the drive shaft4. The electric actuator is an actuator for driving a rotation of the control shaft24.

As shown inFIGS. 1,2and4A-4C, the control shaft24includes a control pivot shaft24aand a plurality of control eccentric cams25. The plurality of control eccentric cams25are provided to every cylinder, and are arranged on an outer circumference of the control pivot shaft24a. The plurality of control eccentric cams25function as a swing fulcrum of the rocker arm15.

The control pivot shaft24aincludes concave portions24band24cformed at a location corresponding to the rocker arm15. Each concave portion24b,24cis formed to have two surfaces opposed to each other in the axial direction of drive shaft4through an axial width. Two bolt-insertion holes26aand26bare formed to pass through the control pivot shaft24ain a radial direction of control pivot shaft24a, in an existing range of the concave portions24band24c. That is, each of the bolt-insertion holes26aand26bis formed between the both concave portions24band24c. These bolt-insertion holes26aand26bare provided to have a predetermined distance from each other in the axial direction. Each of the concave portions24band24cis formed to extend in the axial direction of control pivot shaft24a, and a bottom surface of each concave portion24b,24cis formed flat.

The plurality of control eccentric cams25are constituted by a bracket28and the control eccentric shaft29. The bracket28is fixed to the concave portion24bof control shaft24by two bolts27and27. The two bolts27and27are inserted into the two bolt-insertion holes26aand26bfrom the side of concave portion24c. The control eccentric shaft29is fixed to an tip side of the bracket28.

The bracket28is formed by being bent (by means of bending forming) in an angular-U shape as viewed in a direction perpendicular to the axial direction of control pivot shaft24aand parallel to the bottom surface of each concave portion24b,24c. The bracket28includes a rectangular-shaped base portion28aand arm-shaped fixing portions28band28b. The bracket28(the base portion28a) is formed to extend in a longitudinal direction of the concave portion24b. The base portion28ais fitted into the concave portion24b, and thereby, is held by the concave portion24b. The arm-shaped fixing portions28band28bare provided to both end portions of the bracket28relative to a longitudinal direction of bracket28. That is, the arm-shaped fixing portions28band28bprotrude from the both end portions of bracket28in a lower direction ofFIG. 2.

The base portion28ais formed with female threaded holes in both end-portion sides of base portion28arelative to the longitudinal direction. Tip potions of the bolts27and27are screwed respectively into the female threaded holes of base portion28a. Each of the both fixing portions28band28bis formed with a fixing hole28cin a tip portion of the fixing portion28b. Each fixing hole28cpasses through the fixing portion28b, and serves to fasten the control eccentric shaft29. Moreover, since an outer surface of the base portion28ais in contact with the bottom surface of concave portion24b, and respective outer edge surfaces of both fixing portions28band28bare closely in contact with opposed inner surfaces of concave portion24b, i.e., is fitted to and held by the opposed inner surfaces of concave portion24b; an accuracy of positioning is enhanced relative to the longitudinal direction.

(An outer circumferential surface of) the control eccentric shaft29swingably supports the rocker arm15through the support hole15dof tubular base portion15aof rocker arm15. An axial length L of the control eccentric shaft29is set to be approximately equal to a distance between the respective axially-outside surfaces (outer edge surfaces) of the both fixing portions28band28bof bracket28. The control eccentric shaft29is fixed to the both fixing portions28band28b, e.g., by forcibly inserting both end portions of control eccentric shaft29respectively into the fixing holes28cand28c. A shaft center Q of the control eccentric shaft29serves as a swinging fulcrum of the rocker arm15.

As shown inFIG. 2, axially-outside surfaces of the cam main body5aof drive cam5, axially-outside surfaces of the link rod17and axially-outside surfaces of the swing cam7exist within a range of the length L of control eccentric shaft29, as viewed in a direction perpendicular to the axial direction of drive shaft4.

As shown inFIGS. 4A to 4C, the shaft center Q of control eccentric shaft29is eccentric to (deviated from) a shaft center P of the control pivot shaft24aby a relatively large eccentric amount a because of an arm length of each fixing portion28bof bracket28. In other words, the control eccentric shaft29is formed in a crank shape by use of the bracket28relative to the shaft center P of control pivot shaft24a. Hence, the eccentric amount α can be set at a sufficiently large value.

The electric actuator includes an electric motor and a speed reducer (not shown). The electric motor is fixed to a rear end portion of the cylinder head1. The speed reducer is, for example, a ball screw mechanism for transmitting a rotational drive force of the electric motor to the control pivot shaft24a.

The electric motor is a proportional DC motor. This electric motor is driven by control signals that are outputted from the electronic controller49configured to detect the operating state of engine.

The electronic controller49detects the current operating state of engine, e.g., by calculations using the above-mentioned crank angle sensor for sensing the engine rotational speed, the air flow meter for sensing an amount of intake air, the water-temperature sensor for sensing a water temperature of the engine or the like. Moreover, the electronic controller49detects an operational position of the variable mechanism by receiving information signals derived from a potentiometer for sensing a rotational position of the control shaft24, and the like. Thereby, the electronic controller49controls the electric motor by way of feedback control. Since such an electric actuator uses electricity, a prompt responsivity in change can be obtained irrespective of oil temperature of engine and the like.

The electric actuator controls the valve lift-amount characteristic and the working angle of the intake valve3acontinuously within a range from a minimum value of working angle to a maximum value of working angle, by controlling the rotational position of control pivot shaft24ain accordance with the operating state of engine. That is, a positional relation among the shaft center P of control pivot shaft24a, a shaft center R of the shaft portion15eof rocker arm15, a shaft center S of the pivotally-supporting pin19and the like is assigned (determined) in accordance with the rotational position of control pivot shaft24a. Thereby, an opening timing of valve-lift characteristic is varied toward an advanced side when controlling the midpoint of working angle.

[Operations of Valve Control Apparatus in First Embodiment]

Operations of the valve control apparatus according to the first embodiment will now be explained referring toFIGS. 4A to 9C.FIGS. 4A to 5Cshow a state where the intake valve has been controlled to have a minimum lift amount L1(a minimum working angle D1), by the valve control apparatus.FIGS. 4A to 4Cshow attitudes when the intake valve is closed, andFIGS. 5A to 5Cshow attitudes when the intake valve is open.FIGS. 6A to 7Cshow a state where the intake valve has been controlled to have a middle lift amount L2(a middle working angle D2), by the valve control apparatus.FIGS. 6A to 6Cshow attitudes when the intake valve is closed, andFIGS. 7A to 7Cshow attitudes when the intake valve is open.FIGS. 8A to 9Cshow a state where the intake valve has been controlled to have a maximum lift amount L3(a maximum working angle D3), by the valve control apparatus.FIGS. 8A to 8Cshow attitudes when the intake valve is closed, andFIGS. 9A to 9Cshow attitudes when the intake valve is open.

At first, for example, at the time of low rotational speed of the engine such as idling operation or at the time of low load of the engine, the connection changeover mechanism36does not connect the second swing arm31with the first swing arm30in each cylinder. That is, the electronic controller49does not output the control signal to the electromagnetic changeover valve48, so that the hydraulic-pressure supply/discharge passage43communicates with (i.e., is open to) the drain passage47and does not communicate with (i.e., is closed to) the supply passage46. Hence, hydraulic pressure is not supplied to the pressure-receiving chamber40. As shown inFIG. 2, whole of the connecting pin38is maintained at its backward position by spring force of the coil spring39. That is, the connecting pin38is held within the first retaining hole37aby the biasing force of the coil spring39. Thereby, the first swing arm30is not interlocked with the second swing arm31. Under this state, when the second drive cam13is lifting the second swing arm31, the sip surface35aof the slip convex portion35is in contact with the outer circumferential surface13aof the second drive cam13, so that the shim33bof the second swing arm31is in contact with the stem end of the second intake valve3bby the spring force of the valve spring10b.

At this time, because of the output of control signal from the electronic controller49to the electric motor, the control pivot shaft24ahas been rotated to a counterclockwise-directional position θ1by the ball screw mechanism, as shown inFIGS. 4A to 5C. Hence, the control eccentric shaft29has reached its position corresponding to the position θ1. The shaft center Q has moved away from the drive shaft4in an upper left direction ofFIG. 4A. Thereby, whole of the transmission mechanism8has tilted around the drive shaft4in a counterclockwise direction. Hence, also the swing cam7has rotated in the counterclockwise direction so that a base-circular-surface side of the cam surface7dis in contact with the roller34of the first swing arm30.

When the rocker arm15is raised upwardly by the link arm16in response to the rotation of the drive cam5from the valve-closed state shown byFIG. 4A, the connecting portion7cof swing cam7is lifted upwardly by the link rod17to rotate the swing cam7in the clockwise direction, as shown inFIG. 5A. This lift is transmitted through the roller34of the first swing arm30to the first intake valve3a. Accordingly, the first intake valve3ais lifted and then opened. However, at this time, both of the lift amount and working angle of the first intake valve3aare sufficiently small. (minimum lift amount L1, minimum working angle D1)

On the other hand, the sip surface35aof the second swing arm31is constantly in contact with the outer circumferential surface13aof the second drive cam13. Hence, as shown inFIG. 4C, the second intake valve3bbecomes in the non-lifted state (closed state) when the rotational position of the second drive cam13falls within a base circle region over which the base circular portion of the second drive cam13is in contact with the slip convex portion35. Then, when the rotational position of the second drive cam13falls within a lifted region over which the cam nose portion13bof the second drive cam13is in contact with the slip convex portion35, the second intake valve3bbecomes in the lifted state (open state) as shown inFIG. 5C. In such a low rotational speed or low load state of the engine, the second intake valve3battains a fixed lift curve having a peak lift amount equal to LN and a working angle equal to DN as shown inFIG. 10.

That is, during this control (during the low rotational speed or low load state of the engine), the lift curve L1is realized by the first intake valve3a, and the fixed lift curve LN is realized by the second intake valve3b. As shown inFIG. 10, the peak lift amount LN of the second intake valve3bis smaller than the minimum lift amount L1of the first intake valve3a. Also, the working angle DN of the second intake valve3bis smaller than the minimum working angle D1of the first intake valve3a.

A peak lift phase θN of the second intake valve3bis not deviated much from a peak lift phase θ1of the first intake valve3a, i.e., is substantially equal to the peak lift phase θ1. Accordingly, the lift curve LN is completely accommodated in (i.e., completely lower than) the lift curve L1as shown inFIG. 10. As a result, if the connection changeover mechanism36connects the second swing arm31with the first swing arm30to lift the first and second intake valves3aand3bwith an identical lift characteristic, these intake valves3aand3bare lifted reliably in dependence upon the lift curve L1(by the first drive cam5). In other words, in this case, the common lift characteristic of the connected intake valves3aand3bis not changed from the lift curve L1(that is performed by the first drive cam5) to the lift curve LN (that is performed by the second drive cam13) during the lift operation.

Therefore, a noise generation can be avoided. Moreover, since the lift amount (LN) and the working angle (DN) of the second intake valve3bare respectively smaller than the minimum lift amount (L1) and the minimum working angle (D1) in the control range of the first intake valve3a, the minimum lift amount (L1) and the minimum working angle (D1) of the first intake valve3awhich are necessary for a certain gas exchange (a certain intake-air quantity) can be made relatively large. As a result, variation widths (L1˜L3, D1˜D3) of the lift amount and working angle of the first intake valve3acan be made small. Thereby, an attitude variation of the control mechanism9can be reduced. Hence, a mountability to the engine can be improved. Moreover, a tight attitude (improper attitude) of the control mechanism9can be avoided, resulting in an enhancement in wear resistance of the control mechanism9.

Next, a case where the state of engine has changed to a middle rotational speed region and/or a partial load region because of a steady-state running and the like of the vehicle will now be explained. In such a case, the connection changeover mechanism36still does not connect the second swing arm31with the first swing arm30in each cylinder.

In this case, the control shaft24has further rotated in the counterclockwise direction up to its position θ2by the electric actuator on the basis of the control signal derived from the electronic controller49as shown inFIGS. 6A to 7C. Also, the control eccentric shaft29has rotated up to its position θ2. Thereby, the shaft center Q2of the control eccentric shaft29has become closest (nearest) to the drive shaft4.

Accordingly, whole of the transmission mechanism8including the rocker arm15, the link arm16and the like has rotated around the drive shaft4in the clockwise direction. Hence, also the swing cam7has rotated relatively in the clockwise direction (lifting direction).

In this case, under a state shown byFIGS. 6A to 6C, the base circular surface of the swing cam7is in contact with the roller34so that the cam nose portion7bfaces in the upward direction (toward the control shaft24). Hence, the first intake valve3ais not lifted (i.e., in the closed state). Also the second intake valve3bis not lifted (i.e., in the closed state), because the sip surface35ais in contact with the base circular portion of the second drive cam13so that the cam nose portion13bfaces in the upward direction (toward the control shaft24).

Then, as shown byFIGS. 7A to 7C, a movement of the cam nose portion7bof the drive cam7is transmitted through the first swing arm30to the first intake valve3a. Thereby, the first intake valve3ais lifted. Thus, in the middle load region or the middle rotational speed region of the engine, the valve lift amount and the working angle of the first intake valve3aare increased as shown inFIG. 10. Therefore, in this engine region, the middle lift amount L2and the middle working angle D2of the first intake valve3aare obtained.

At this time, the cam nose portion13bof the second drive cam13downwardly presses the sip surface35aso as to lift and open the second intake valve3b. In this case, the second intake valve3battains the fixed lift curve LN (having the peak lift amount equal to LN) as shown inFIG. 10. At a drive-shaft angle at which the first intake valve3atakes its peak lift, the second intake valve3btakes a lift amount value somewhat smaller than the peak lift amount LN, as shown inFIG. 10. In other words, a peak-lift phase of the first intake valve3ais slightly retarded as compared with a peak-lift phase of the second intake valve3b.

Next, a case where the state of engine has changed to a high rotational speed region or a high load region will now be explained. In such a case, the electromagnetic changeover valve48communicates the hydraulic-pressure supply/discharge passage43with the supply passage46and blocks the communication between the hydraulic-pressure supply/discharge passage43and the drain passage47, by the signal outputted from the electronic controller49. Thereby, high-pressure oil is supplied to the pressure-receiving chamber40, so that the front-end portion38aof the connecting pin38is inserted into the second retaining hole37bso as to engage with the first swing arm30when the first swing arm30is not being lifted.

That is, at this time, the second swing arm31is in non-lifted state. Hence, when the first swing arm30is also in the non-lifted state, the first retaining hole37aconforms to the second retaining hole37b. Therefore, when both of the first and second swing arms30and31are in the non-lifted state, the connecting pin38moves in the right direction ofFIG. 2against the biasing force of coil spring39so that the front-end portion38aenters the second retaining hole37bto be engaged. Accordingly, the first swing arm30is integrally connected (interlocked) with the second swing arm31, so that the first swing arm30repeats the lifting operation and its returning operation in synchronization with the second swing arm31.

Under this case, the control pivot shaft24ahas rotated in the counterclockwise direction up to a position θ3by the ball screw mechanism because the control signal has been outputted from the electronic controller49to the electric motor, as shown inFIGS. 8A to 9C. Hence, the control eccentric shaft29has reached its position corresponding to the position θ3. The shaft center Q has moved away from the drive shaft4in an upper right direction ofFIG. 8A. Thereby, whole of the transmission mechanism8has tilted around the drive shaft4in the clockwise direction. Hence, also the swing cam7has rotated in the clockwise direction around the drive shaft4, so that the contact point between the cam surface7dand the roller34of the first swing arm30has approached a lift-surface side of cam surface7d.

FIGS. 8A to 8Cshow attitudes of this case under the non-lifted state corresponding to the valve-closed state. As shown inFIG. 8A, the base circular surface of the swing cam7is in contact with the roller34so that the cam nose portion7bfaces in the upward direction (toward the control shaft24). Hence, the first intake valve3ais in the not-lifted state (i.e., in the closed state). Also the second intake valve3bis in the not-lifted state (i.e., in the closed state), because the sip surface35ais in contact with the base circular portion of the second drive cam13so that the cam nose portion13bfaces in the upward direction.

FIGS. 9A to 9Cshow attitudes of this case under a state where the first intake valve3ais open. That is,FIGS. 9A to 9Cshow a moment when an eccentric direction Y-X of the first drive cam5(i.e., a direction from the shaft center Y of drive shaft4toward the center X of cam main body5a) has just faced in an axis-distance direction of the link arm16(i.e., a direction from X toward R). At this time, as shown inFIG. 10, the first intake valve3atakes the maximum peak lift amount L3, and realizes the maximum working angle D3.

As mentioned above, the two swing arms30and31operate integrally with each other because the connection changeover mechanism3has already connected the second swing arm31with the first swing arm30. Hence, the second intake valve3btakes the same lift curve as the first intake valve3a. That is, as shown inFIG. 9C, a large clearance C exists between the cam nose portion13bof the second drive cam13and the sip surface35aof the second swing arm31, and hence, the lift (rotation) of the cam nose portion13bof the outer circumferential surface13aof the second drive cam13is not transmitted to the second swing arm31. Accordingly, in the same manner as the first intake valve3a, the second intake valve3btakes the maximum peak lift amount L3and realizes the maximum working angle D3, in dependence upon the swinging motion of the first swing arm30.

Next, advantageous effects in the first embodiment will now be explained from a viewpoint of a performance of the engine.

In the control condition of the minimum lift amount L1(minimum working angle D1) as shown inFIGS. 4A to 5C, the first intake valve3atakes the lift curve L1whereas the second intake valve3btakes the lift curve LN shown inFIG. 10. As mentioned above, this control condition is used in the low rotational-speed region of engine such as idling. Thus, by reducing the lift working angle D, a pumping loss is reduced while a friction is reduced, resulting in an improvement of fuel economy.

Moreover, the second intake valve3bis made to take a lift amount and a working angle as small as possible. Thereby, a lift difference between the first and second intake valves3aand3bis enlarged so that a swirl effect is enhanced to improve a combustion of the engine. Accordingly, the fuel economy can be further improved.

If the lift or the working angle of the second intake valve3bis set to be excessively small, there is the following risk. That is, it is easy for a deposit to adhere to a portion near a contact portion between a valve seat and an outer circumference of an umbrella portion of the second intake valve3bwhen the second intake valve3bis in the closed state. Specifically, a component derived from a reflowed mixture gas (air-fuel mixture) or EGR gas sticks to the portion near the contact portion and grows as the deposit when the second intake valve3bis in the closed state.

In the first embodiment according to the present invention, when the second intake valve3bopens, gas flows to the outer circumference of the umbrella portion at a high speed so that the deposit is broken up and removed.

This advantageous effect in the first embodiment becomes higher as the working angle of the second intake valve3bbecomes larger or as the lift amount of the second intake valve3bbecomes larger. However, if the working angle or lift amount of the second intake valve3bis excessively large, the swirl effect which is caused by the lift difference between the first and second intake valves3aand3bis weak.

Therefore, working angle and lift amount which are the minimum necessary to enable the deposit removal are required. In the first embodiment according to the present invention, the lift curve LN which is performed by the second drive cam13is set at the predetermined fixed lift curve (only one lift curve). This predetermined fixed lift curve satisfies the deposit-removal requirement and also produces a sufficient swirl effect. Moreover, this lift curve LN for the second intake valve3bdoes not vary even if the working angle or the peak lift amount of the first intake valve3avaries. That is, the deposit removal and the enhancement of swirl can be stably maintained irrespective of the variation of the working angle or peak lift amount of the first intake valve3a.

For example, in the control condition of the middle lift amount L2(middle working angle D2) where the swing arms30and31have been not connected with each other as shown inFIGS. 6A to 7C, the second intake valve3aperforms a lift curve substantially identical with the lift curve LN. Also in this control condition, the deposit removal and the enhancement of swirl can be stably maintained.

In this control condition, i.e., in the partial-load region over which the load (or rotational speed) is higher than that of the idling operation, fuel consumption can be reduced by virtue of a combustion improvement obtained by the swirl effect.

In the operating condition that a required torque is high, an opening of a throttle valve (not shown) is increased. At the same time, the connection changeover mechanism36connects the second swing arm31with the first swing arm30as shown inFIGS. 8A to 9C. As a result, both of the first and second intake valves3aand3bare controlled with the maximum lift amount L3(the maximum working angle D3). Thereby, the intake air quantity is increased, so that the torque (output) can be enhanced. Thus, the intake air quantity is increased in the high-torque region, and thereby, the combustion is improved. Therefore, in this condition, the swirl effect is not necessary.

As shown by a lift characteristic view in a right side ofFIG. 11, in the case where the first and second swing arms30and31are in the connected (interlocked) state by the connection changeover mechanism36, both of the first and second intake valves3aand3brealize the same lift curve. In such a case, the common working angle of both the first and second intake valves3aand3bvaries from the working angle D1of the lift curve L1having the peak lift amount L1to the working angle D3of the lift curve L3having the peak lift amount L3. A maximum output power may be enhanced by making the working angle larger as the engine rotational speed becomes higher, and a very-low-rotation torque may be enhanced by making the working angle narrower as the engine rotational speed becomes lower.

FIG. 12shows one example of a control map for the peak lift amounts of the first and second intake valves3aand3b.

The map ofFIG. 12has an X-axis of the engine rotational speed and a Y-axis of the engine torque (load). In a case that the torque is lower than a K-line of this map, the connection changeover mechanism36disconnects the second swing arm31from the first swing arm30so as to keep the lift difference between the first and second intake valves3aand3b. Accordingly, the combustion is improved by the swirl effect, resulting in the improvement of fuel economy.

On the other hand, in a case that the torque is higher than the K-line on the map ofFIG. 12, the connection changeover mechanism36connects the second swing arm31with the first swing arm30so as to lift both the first and second intake valves3aand3bwith a relatively large lift amount. Accordingly, the torque is increased.

As shown inFIG. 12, a torque (Y-axis) of the K-line decreases with the rise of the engine rotational speed (X-axis). That is, the connection changeover mechanism36connects the first and second swing arms30and31with each other in advance at the time of a lower torque as the engine rotational speed becomes higher, because a frequency at which the vehicle runs with high torque becomes higher as the engine rotational speed becomes higher. Thereby, the number of times the connection changeover mechanism36connects/disconnects the second swing arm31with/from the first swing arm30is reduced, and moreover, a frequency at which a time delay necessary for the connection/disconnection (i.e., switching) of the swing arms30and31occurs can be reduced. Accordingly, a smooth torque rise can be attained. Also, a frequency at which a torque shock occurs due to the connecting/disconnecting operation (switching operation) of the connection changeover mechanism36can be lowered.

If the lift amount of the second intake valve3bis changed rapidly from the very-small lift LN to the large lift equal to that of the first intake valve3awhen an operating point of the engine exceeds the K-line, the above-mentioned torque shock occurs due to the rapid torque rise. Therefore, in the first embodiment according to the present invention, a transient lift control is performed as shown inFIG. 13.

FIG. 13shows an example in which the vehicle accelerates from the idling. This example is also shown by a thick line ofFIG. 12. A solid line ofFIG. 13represents a variation characteristic of the peak lift amount of the first intake valve3a. A dotted line ofFIG. 13represents a variation characteristic of the peak lift amount of the second intake valve3b. At first, the second intake valve3btakes the very-small fixed peak lift LN whereas the first intake valve3atakes the peak lift L1. Then, the first intake valve3agradually increases its peak lift amount with the increase of engine speed and the increase of engine load. Then, the operating point of the engine reaches the K-line at which the peak lift of the first intake valve3areaches the middle peak lift L2. At this time (on the K-line), if the connection changeover mechanism36connects the second swing arm31with the first swing arm30, the peak lift of the second intake valve3bsharply rises from the very-small lift LN to the middle lift L2so that the air quantity is also rapidly increased. In this case, there is a risk that the torque rises sharply to cause the torque shock.

Therefore, in the first embodiment according to the present invention, concurrently when the connection changeover mechanism36connects the second swing arm31with the first swing arm30, the common peak lift amount for the both intake valves3aand3bis changed from the lift amount L2to a lift amount L1.5as shown inFIG. 13by rotating the control shaft24in one direction.

Thus, when the connection changeover mechanism36connects the second swing arm31with the first swing arm30, both of the first and second intake valves3aand3bare made to take the valve lift amount L1.5. The valve lift amount L1.5which is realized by both the first and second intake valves3aand3bproduces a total torque substantially equal to that produced when the first intake valve3atook the valve lift amount L2and the second intake valve3btook the valve lift amount LN. Hence, the torque shock due to torque level-difference as mentioned above is reduced or suppressed.

In the first embodiment, the example has been explained in which the valve control apparatus according to the present invention is applied to the first and second intake valves3aand3b. However, the valve control apparatus according to the present invention can be applied also to first and second exhaust valves.

That is, it is easy for a deposit of combustion gas to adhere to a portion near a contact portion between a valve seat and an outer circumference of an umbrella portion of the second exhaust valve when the second exhaust valve is in the closed state. This deposit can be removed by setting the lift characteristic of the second exhaust valve at the fixed very-small lift (curve) LN. Even if the lift amount characteristic of the first exhaust valve is varied, this lift curve LN for the second exhaust valve is not varied. Thereby, the deposit can be reliably removed.

Since the very-low lift LN of the second exhaust valve is maintained, combustion gas is mainly exhausted from the first exhaust valve. During an exhaust stroke, a gas flowing is strengthened within the cylinder so that a combustion stability in next combustion cycle is improved. Accordingly, the fuel consumption can be reduced. Moreover, since an exhaust gas flow to a downstream exhaust manifold and a catalyst is disturbed, a conversion performance of the catalyst is enhanced so that an exhaust emission can be reduced.

FIGS. 14 to 17show a second embodiment according to the present invention. In the second embodiment, each of the first drive cam5and a second drive cam50is formed integrally with the drive shaft4. Moreover, the swing cam7including the cam shaft7ais formed such that the swing cam7can be divided (separated) into two pieces via its base end portion (located between the connecting portion7cand the cam nose portion7b). Hence, the cam shaft7aof the swing cam7is also dividable.

That is, both of the first drive cam5and the second drive cam50are formed integrally with the drive shaft4when the drive shaft is molded by casting, forging or the like. This second drive cam50is formed as a large oval cam (large egg-shaped cam) as compared with the second drive cam13of the first embodiment.

Because the first and second drive cams5and50are molded integrally with the drive shaft4as mentioned above, the drive shaft4cannot be inserted sequentially into the plurality of swing cams7from the end portion of the drive shaft4due to the existence of the drive cams5and50when trying to mount the swing cams7on the drive shaft4. Hence, the swing cam7which has the shape of the first embodiment cannot be attached to the drive shaft4of the second embodiment.

Therefore, in the second embodiment, as shown inFIG. 14, the swing cam7is formed as two separate pieces of a cam main body and a bracket member7e. These cam main body and the bracket member7eare dividable at the base end portion side of the swing cam7(located between the connecting portion7cand the cam nose portion7b). The cam main body has the cam surface7d. Each of these cam main body and bracket member7eincludes a bearing groove formed in a half-round shape. The bearing grooves are fitted over the drive shaft4from a radially outside of the drive shaft4so as to face each other, and under this state, the bracket member7eis combined with the cam main body by using two bolts14and14.

As mentioned above, since the first and second drive cams5and50are provided integrally with the drive shaft4, a support stiffness of each of the first and second drive cams5and50becomes high so that a lift behavior can be stabilized. Moreover, because the fixing pin12as mentioned in the first embodiment is unnecessary, the number of components and the cost of manufacturing can be reduced.

Moreover, as shown inFIGS. 14 and 15, one end portion of the cam shaft7aof the swing cam7which is located on the side of the first drive cam5is formed to extend in the axial direction. A front edge of this extension portion7fis located near one lateral surface of the first drive cam5. Thus, by providing the extension portion7f, the fall of swing cam7in the axial direction can be suppressed during its swinging motion. Moreover, by removing a sleeve2which is provided in the first embodiment, the number of components can be reduced.

The link rod16is mounted by inserting the drive shaft4into the link rod16in the axial direction, i.e., from the lateral direction.

In the second embodiment, a second roller51is rotatably supported by a second roller shaft51aat a substantially center portion of the second swing arm31relative to a longitudinal direction of the second swing arm31. Hence, an outer circumferential surface50aof the second drive cam50is rotatably in contact with the second roller51, instead of the slip surface of the first embodiment. This structure is given for the purpose of suppressing an increase of friction loss because the second drive cam50is enabled to produce a relatively high lift.

Accordingly, in the second embodiment, for example, under the unconnected state where the connection changeover mechanism36has not yet connected the second swing arm31with the first swing arm30in a predetermined rotational-speed region of the engine, the first roller34is rotatably in contact with the cam surface7dof the swing cam7so as to lift (open) the first intake valve3a. Thereby, the lift amount L and the working angle D of the first intake valve3avary between the lift curve characteristics L1to L3ofFIG. 17. On the other hand, under this state, the second intake valve3balways take a fixed lift curve depending on a cam profile of the second drive cam50. This fixed lift curve is shown by a lift curve LN ofFIG. 17which has a peak lift amount LN and a working angle DN.

Then, when the connection changeover mechanism36connects the first swing arm30with the second swing arm31in a high speed region of the engine or the like, the lifts of the intake valves3aand3bare controlled by the cam profile of the second drive cam50which can produce a large lift, as shown inFIGS. 16A to 16C. Thereby, a clearance C1is given between the cam surface7dof the swing cam7and the first roller34as shown inFIG. 16A, so that the first intake valve3aopens in dependence upon the lift amount of the second drive cam50, together with the second intake valve3b.

That is, as show inFIG. 17, the lift amount LN and the working angle DN of the second intake valve3bare respectively larger than the maximum lift amount L3and the maximum working angle D3of the first intake valve3awhich are controlled by the cam surface7dof the swing cam7. Accordingly, when the connection changeover mechanism36has already connected the first swing arm30with the second swing arm31, both of the first and second intake valves3aand3bare driven by the lift curve LN which is performed by the second drive cam50.

FIG. 18show a summary of the lift characteristics of the first and second intake valves3aand3bin the second embodiment. As seen fromFIG. 18, the second intake valve3bconstantly operates (opens) with the large lift amount LN and the large working angle DN. Accordingly, torque can be increased only by opening the throttle valve (not shown), so that a rising responsivity of torque is enhanced.

Contrary to this, in the case of the first embodiment, when a sudden acceleration is required under a running state where the first intake valve3ais operating with the small working angle L1and the second intake valve3bis operating with the very-small working angle LN, it is necessary to increase the working angle and also connect the first and second swing arms30and31with each other in order to increase torque. By that much, the torque generation needs time.

In the second embodiment, the lift amount LN of the second intake valve3bwhen the connection changeover mechanism36is in the released state is larger than the maximum lift amount L3which is obtainable within the control lift range of the first intake valve3a. Moreover, the working angle DN of the second intake valve3bwhen the connection changeover mechanism36is in the released state is larger than the maximum working angle D3which is obtainable within the control lift range of the first intake valve3a.

Therefore, when the first and second swing arms30and31have been connected with each other by the connection changeover mechanism36, any of the first and second intake valves3aand3bcan be prevented from being partially driven by the first drive cam5during the lifting operation. That is, the drive by the second drive cam50can be prevented from being changed into the drive by the first drive cam5. Hence, noise can be reduced.

Moreover, since the lift amount LN and the working angle DN of the second intake valve3bare larger than the maximum lift amount L3and the maximum working angle D3which are obtainable within the control range of the first intake valve3aby the first drive cam5, the maximum lift amount D3and the maximum working angle D3of the first intake valve3awhich are necessary for a certain gas exchange can be set at relatively small values. As a result, the variation widths (L1˜L3, D1˜D3) of the lift amount and working angle of the first intake valve3acan be made small, so that an attitude change of the transmission mechanism8can be suppressed. Accordingly, the mountability to the engine and the like can be improved. Moreover, the transmission mechanism8can be inhibited from being forced to take a tight attitude (improper attitude), so that wear and abrasion resistance of the transmission mechanism8can be enhanced.

In the second embodiment, the example has been explained in which the valve control apparatus according to the present invention is applied to the intake valves. However, the valve control apparatus according to the present invention can be applied also to exhaust valves. In such a case, peak lift amount and working angle of one of the exhaust valves are varied whereas peak lift amount and working angle of another of the exhaust valves are fixed relative to the load and rotational speed of the engine. These fixed peak lift amount and fixed working angle of the another of the exhaust valves are respectively larger than the peak lift amount and working angle of the one of the exhaust valves. That is, the another of the exhaust valves realizes a fixed lift curve having the fixed peak lift amount and the fixed working angle. Accordingly, in the same manner as the above example in the second embodiment, the noise reduction and the variation-width reduction in lift amount and working angle can be attained.

FIGS. 19A to 19Cshow a third embodiment according to the present invention. A basic structure of the valve control apparatus of the third embodiment is the same as the second embodiment. However, in the third embodiment, the first intake valve3aopens and closes during the exhaust stroke whereas the second intake valve3bopens and closes during an intake stroke as usual. That is, the first drive cam is fixed (fastened) to the drive shaft4at a relatively phase-advanced position. Contrary to this, the second drive cam is fixed to the drive shaft4at a relatively phase-retarded position.

FIGS. 19A to 19Cshow attitudes at a moment when the peak lift of the first intake valve3ajust takes the value L3under the state where the first intake valve3ais being controlled by the lift curve L3in the unconnected state of the connection changeover mechanism36. At this moment, as shown inFIG. 19C, the second intake valve3bis in the non-lifted state (closed state) because the second drive cam50is fixed to the drive shaft4at its position retarded in phase largely by μ in the counterclockwise direction.

Then, when the drive shaft4has just rotated by μ in phase, the second intake valve3btakes the peak lift amount LN by means of the second drive cam50. Hence, as shown inFIG. 20and a left part of21, the fixed lift curve LN of the second intake valve3bstarts (i.e., has positive values) after the lift curve L3of the first intake valve3aends (i.e., becomes zero).

The lift curve L3of the first intake valve3amay be set to be included in (i.e., to be entirely smaller than) a lift curve of each of two exhaust valves provided in every cylinder. This lift curve of each exhaust valve is shown by a dotted line inFIG. 20orFIG. 21. In such a case, the lift (opening action) of the first intake valve3astarts after a lift (opening action) of each exhaust valve started. Then, the lift (open state) of the first intake valve3aends before the lift (open state) of each exhaust valve ends. Therefore, exhaust gas (EGR gas) can be prevented from flowing at high pressure back to the intake side to cause a suction noise.

In the third embodiment, the minimum lift curve L1of the first intake valve3ais set to be constantly equal to 0, i.e., is set not to lift the first intake valve3a. This minimum lift curve L1can be easily set by changing the position in phase of the control shaft24in the more clockwise direction inFIGS. 19A to 19C, or alternatively by causing a cam protruding shape of the swing cam7to be lower than that of the first embodiment.

Next, when the connection changeover mechanism36has connected the second swing arm31with the first swing arm30, both of the first and second intake valves3aand3bperform a sub lift during the exhaust stroke and then perform a main lift according to the fixed lift curve LN during the intake stroke, as shown by a right part ofFIG. 21.

Since both of the intake valves3aand3bare opened, an intake-air charging efficiency is enlarged resulting in torque increase. Particularly, if torque is required to increase at the utmost extent, the sub lift is set to take the lift curve L1, i.e., is set to produce no lift. In this case, an EGR amount introduced into the cylinder is minimized, so that a charging efficiency of fresh air is enhanced to increase the torque to the utmost extent. If torque is not required to increase so much, the sub lift is set to take some actual lift to introduce some degree of EGR amount. Thereby, the fuel economy can be improved.

A summary of engine-performance effects under the state where the connection changeover mechanism36is in the non-connected state in the third embodiment is as follows. That is, during the exhaust stroke, the working angle and lift amount of the first intake valve3awhich performs the sub lift are controllably varied, and thereby, the gas amount of EGR which is discharged toward the intake port can be adjusted. At this time, the EGR gas is discharged only from the first intake valve3a, but is not discharged from the second intake valve3b, so that a swirl within the cylinder occurs during the exhaust stroke.

Moreover, since the lift characteristic of the second intake valve3bwhich performs the main lift during next intake stroke is the fixed one, a stable air-intake operation can be achieved even if the characteristic of the sub lift is controllably varied. Additionally, since this main lift is done only by the second intake valve3b, the swirl occurs also during the intake stroke.

By virtue of the above-mentioned EGR gas-amount adjustment, the exhaust-stroke swirl, the intake-stroke swirl, the stabilization of air-intake operation, and the like; the engine performance such as the fuel economy and an exhaust performance can be improved.

Moreover, by virtue of these, a permissible value of the gas amount of EGR which is introduced into the cylinder can be enlarged. Also from this point of view, the fuel economy and the exhaust performance can be further improved.

On the other hand, as the engine-performance effects under the state where the connection changeover mechanism36is in the connected state in the third embodiment, for example, the intake-air charging efficiency can be increased to increase the torque because both the intake valves3aand3bare opened (lifted) as mentioned above.

FIGS. 22 and 23show a fourth embodiment according to the present invention. A basic structure of the valve control apparatus in the fourth embodiment is the same as the third embodiment. However, in the fourth embodiment, the valve control apparatus is applied to the exhaust valves in place of the intake valves. That is, as different points from the third embodiment, the first intake valve3aof the third embodiment is replaced with a first exhaust valve3a, and the second intake valve3bof the third embodiment is replaced with a second exhaust valve3b. Moreover, the phase of the second drive cam50is advanced by μ in the fourth embodiment although the phase of the second drive cam50is retarded by μ in the third embodiment.

As a result, as shown inFIG. 23, after a main lift action of the second exhaust valve3bis performed with the fixed lift curve LN during the exhaust stroke, a sub lift action of the first exhaust valve3ais performed during the intake stroke.

In the fourth embodiment, each of first and second intake valves (not shown) realizes a fixed large lift curve LI (large lift amount) as shown by dotted lines ofFIGS. 22 and 23.

A maximum sub lift curve L3of the first exhaust valve3amay be set to be included in (i.e., to be entirely smaller than) the lift curve LI of the two intake valves. In this case, the exhaust valve opens after the intake valve opens, and then, the exhaust valve closes before the intake valve closes. Hence, the exhaust gas (EGR gas) is inhibited from entering the cylinder under high pressure to heat the inside of cylinder. Therefore, an induction of knocking can be suppressed.

The minimum lift curve L1of the first exhaust valve3ais set to produce no lift (i.e., is set to have no opening time).

Next, when the connection changeover mechanism36has connected the second swing arm31with the first swing arm30, the two exhaust valves3aand3bperform a sub lift during the intake stroke and then perform a main lift according to the fixed lift curve LN during the exhaust stroke subsequent to the combustion, as shown by a right part ofFIG. 23. Since both of the exhaust valves3aand3bare opened during the exhaust stroke, an exhaust efficiency is enlarged resulting in torque increase.

Particularly, if torque is required to increase at the utmost extent, the sub lift is set to take the lift curve L1, i.e., is set to produce no lift. In this case, an EGR amount introduced into the cylinder is minimized during the intake stroke, so that the charging efficiency of fresh air is enhanced to increase the torque to the utmost extent. If torque is not required to increase so much, the sub lift is set to take some actual lift to introduce some degree of EGR amount. Thereby, the fuel economy can be improved.

A summary of engine-performance effects under the state where the connection changeover mechanism36is in the non-connected state in the fourth embodiment is as follows. That is, during the intake stroke, the working angle and lift amount of the first exhaust valve3awhich performs the sub lift action are controllably varied, and thereby, the gas amount of EGR which flows from the exhaust port side into the cylinder can be adjusted. At this time, the EGR gas flows in only from the first exhaust valve3a, but does not flow in from the second exhaust valve3b, so that a swirl within the cylinder occurs during the intake stroke.

Moreover, since the lift characteristic of the second exhaust valve3bwhich performs the main lift during next exhaust stroke subsequent to combustion is the fixed one, a stable exhaust operation can be achieved even if the characteristic of the sub lift is controllably varied. Additionally, since this main lift is done only by the second exhaust valve3b, the swirl occurs also during the exhaust stroke. A part of this swirl remains during next intake stroke, so that the above-mentioned swirl during the intake stroke can be further enhanced.

By virtue of the above-mentioned EGR gas-amount adjustment, the exhaust-stroke swirl, the intake-stroke swirl, the stabilization of exhaust operation, and the like; the engine performance such as the fuel economy and the exhaust performance can be improved.

Moreover, by virtue of these, a permissible value of the gas amount of EGR which is introduced into the cylinder can be enlarged. Also from this point of view, the fuel economy and the exhaust performance can be further improved.

On the other hand, as the engine-performance effects under the state where the connection changeover mechanism36is in the connected state in the fourth embodiment, for example, the exhaust efficiency can be increased to increase the torque because both the exhaust valves3aand3bare opened (lifted) during the exhaust stroke as mentioned above.

Although the present invention has been described above with reference to the embodiments of the present invention, the present invention is not limited to the embodiments described above. Modifications and variations of the embodiments described above will occur to those skilled in the art in light of the above teachings.

In the above respective embodiments, the pair of swing arms30and31which are configured to swing about the rocker shaft32are provided as the pair of followers. Moreover, the connection changeover mechanism36is provided between the pair of swing arms30and31. However, according to the present invention, the pair of swing arms30and31may be replaced with another-type ones, as the pair of followers. For example, a pair of cylindrical valve lifters of direct-acting type may be provided such that the pair of engine valves are driven respectively via the pair of cylindrical valve lifters of direct-acting-type.

A part of lateral surface of cylindrical shape of each of the valve lifters may be formed with a flat surface portion such that a connection changeover mechanism is provided between the flat surface portions which are in contact with each other.

In the above respective embodiments, the connection changeover mechanism36is constructed by the connecting pin38. However, according to the present invention, the connection changeover mechanism is not limited to this structure. The connection changeover mechanism may be of prop type (lever type) as shown in Japanese Patent Application Publication No. H08-210113. Moreover, the drive source for the connecting pin is not limited to the hydraulic pressure (oil pressure). That is, according to the present invention, the connecting pin may be driven by an electromagnetic solenoid as shown in Japanese Patent Application Publication No. 2012-002095.

Moreover, in the above respective embodiments, the variable mechanism which continuously varies the lift amount of the first engine valve and thereby operates the first engine valve is driven by the eccentric cam provided as the drive cam. However, according to the present invention, the drive cam is not limited to the eccentric cam, but may be an egg-shaped cam as shown in Japanese Patent Application Publication No. 2007-321653 (corresponding to US Patent Application Publication No. 2007/0277755).

Moreover, a variable mechanism which can vary the phase may be provided together with a chain sprocket (not shown) provided at a tip portion of the drive shaft, as shown in Japanese Patent Application Publication No. 2009-074414 (corresponding to US Patent Application Publication No. 2009/0078223). In such a case, a correlation between intake valve timing and exhaust valve timing can be varied, so that a further improvement of performance is promising.

Some technical configurations obtainable from the above embodiments according to the present invention will now be listed with their advantageous effects.

[a] A valve control apparatus for an internal combustion engine, comprising: a first engine valve (3a) biased in a closing direction of the first valve (3a) by a biasing force of a valve spring (10a); a second engine valve (3b) biased in a closing direction of the second valve (3b) by a biasing force of a valve spring (10b); a first drive cam (5) provided on a drive shaft (4) and configured to rotate integrally with the drive shaft (4), the drive shaft (4) being configured to rotate in synchronization with a crankshaft; a second drive cam (13,50) provided on the drive shaft (4) and configured to rotate integrally with the drive shaft (4); a swing cam (7) configured to swing; a transmission mechanism (8) configured to convert a rotational motion of the first drive cam (5) into a swinging force and to transmit the swinging force to the swing cam (7); a first swing arm (30) configured to open the first engine valve (3a) by being pressed by a swing of the swing cam (7); a second swing arm (31) configured to open the second engine valve (3b) by being pressed by a rotation of the second drive cam (13,50); a control mechanism (9) configured to vary a swing amount of the swing cam (7) by varying an attitude of the transmission mechanism (8); and a connection changeover mechanism (36) configured to connect and disconnect the first swing arm (30) with/from the second swing arm (31). Accordingly, when the connection changeover mechanism (36) has disconnected the first swing arm (30) from the second swing arm (31), a lift amount characteristic of one of the engine valves (3a,3b) does not vary in conjunction with a lift amount characteristic of another of the engine valves (3a,3b) because both the swing arms (30,31) are not influenced from each other.

[b] The valve control apparatus as described in the above item [a], wherein the first and second engine valves (3a,3b) are first and second intake valves, and a lift characteristic of the second intake valve is set to have a predetermined lift amount (LN) and a predetermined working angle (DN) which are smaller than a minimum lift amount and a minimum working angle obtainable within a control range of the first intake valve, in a case that the connection changeover mechanism (36) has disconnected the first swing arm (30) from the second swing arm (31).

[c] The valve control apparatus as described in the above item [b], wherein an outer diameter of the second drive cam (13) is smaller than an outer diameter of the drive shaft (4).

[d] The valve control apparatus as described in the above item [b], wherein the first swing arm (30) includes a roller (34) rotatably abutting on the swing cam (7).

Since the swing cam (7) changes its frictional direction at the contact portion between the first swing arm (30) and the swing cam (7), the swing cam (7) is easy to wear. However, by using such a roller (34), the generation of wear (abrasion) can be suppressed.

[e] The valve control apparatus as described in the above item [b], wherein the second swing arm (31) includes a contact surface (35a) configured to become in contact with the second drive cam (13).

Since a frictional direction of the rotating second drive cam (13) is fixed (not changed), the contact portion between the second drive cam (13) and the second swing arm (31) is difficult to wear. Hence, the contact portion between the second drive cam (13) and the second swing arm (31) can be constituted by a mere contact surface (35a) without a roller. Accordingly, the cost reduction can be attained as compared with the case that a roller is provided.

[f] The valve control apparatus as described in the above item [a], wherein the connection changeover mechanism (36) includes a connection hole (37b) formed in the first swing arm (30), a connection hole (37a) formed in the second swing arm (31), a connecting member (38) provided to be able to move inside the connection holes (37a,37b) of the first and second swing arms (30,31), a biasing member (39) provided in at least one of the connection holes (37a,37b) of the first and second swing arms (30,31), and configured to bias the connecting member (38) in one direction, and a hydraulic-pressure supply passage (43) through which a hydraulic pressure for moving the connecting member (38) against a biasing force of the biasing member (39) is supplied to at least one of the connection holes (37a,37b) of the first and second swing arms (30,31).

[g] The valve control apparatus as described in the above item [a], wherein a characteristic of the second engine valve is set to have a predetermined lift amount (LN) which is larger than a maximum lift amount obtainable within a control range of the first engine valve and to have a predetermined working angle (DN) which is larger than a maximum working angle obtainable within the control range of the first engine valve, in a case that the connection changeover mechanism (36) is in a non-connected state.

[h] The valve control apparatus as described in the above item [g], wherein the first swing arm (30) is equipped with a roller (34) configured to freely rotate at a contact portion between the swing cam (7) and the first swing arm (30), and the second swing arm (31) is equipped with a roller (51) configured to freely rotate at a contact portion between the second drive cam (50) and the second swing arm (31)

Accordingly, a stable swing can be attained by the rotatable contact by use of the roller (34) in the case that the fixed lift is large.

[i] The valve control apparatus as described in the above item [h], wherein the swing cam (7) is constituted by dividable two members that sandwich the drive shaft (4) therebetween.

Accordingly, the swing cam (7) can be attached, for example, even if the second drive cam is formed integrally with the drive shaft (4). Hence, an assembling workability is improved.

[j] The valve control apparatus as described in the above item [a], wherein the first and second engine valves (3a,3b) are first and second intake valves, and opening and closing of the first intake valve are performed during an exhaust stroke, and opening and closing of the second intake valve are performed during an intake stroke, in a case that the connection changeover mechanism (36) is in a non-connected state.

Accordingly, a suction of EGR gas can be conducted because one of the intake valves is opened during the exhaust stroke. Therefore, the fuel economy is improved. Moreover, a swirl of the EGR gas can be produced because only one of the intake valves is lifted.

[k] The valve control apparatus as described in the above item [j], wherein an open period of the first intake valve does not overlap with an open period of the second intake valve in the case that the connection changeover mechanism (36) is in the non-connected state.

Accordingly, a stable operation can be realized because a drive cam which is actually opening the two valves when the connection changeover mechanism (36) is in the connected state is not switched between the two drive cams during the open state of the valves.

[l] The valve control apparatus as described in the above item [k], wherein working angle and lift amount of the first intake valve are smaller than working angle and lift amount of an exhaust valve even when each of the working angle and lift amount of the first intake valve takes a maximum level obtainable within a control range thereof.

Accordingly, excessive amount of exhaust gas can be inhibited from being introduced into the intake port on the side of the intake valve because the intake valve opens and closes within a range of the lift amount of the exhaust valve. As a result, a problem that the exhaust gas hits against an air cleaner and the like to generate an abnormal noise can be suppressed.

[m] The valve control apparatus as described in the above item [k], wherein a swing amount of the first swing arm (30) which is derived from the swing cam (7) is substantially equal to 0 in a case that the connection changeover mechanism (36) is in a connected state.

That is, the first swing arm does not open the valve during the exhaust stroke when the connection changeover mechanism is in the connected state. Thereby, a rate of fresh air is increased so that torque can be increased in the high speed region or the like of the engine in which high torque is needed.

[n] The valve control apparatus as described in the above item [a], wherein the first and second engine valves (3a,3b) are first and second exhaust valves, and opening and closing of the first exhaust valve are performed during an intake stroke, and opening and closing of the second exhaust valve are performed during an exhaust stroke, in a case that the connection changeover mechanism (36) is in a non-connected state.

Accordingly, a suction of EGR gas can be conducted because one of the exhaust valves is opened (lifted) during the intake stroke. Therefore, the fuel economy is improved. Moreover, a swirl of the EGR gas can be produced because only one of the exhaust valves is lifted.

[o] The valve control apparatus as described in the above item [n], wherein an open period of the first exhaust valve does not overlap with an open period of the second exhaust valve in the case that the connection changeover mechanism (36) is in the non-connected state.

Accordingly, a stable operation can be realized because a drive cam which is actually opening the two valves when the connection changeover mechanism (36) is in the connected state is not switched between the two drive cams during the open state of the valves.

[p] The valve control apparatus as described in the above item [o], wherein the open period and lift amount of the first exhaust valve are smaller than open period and lift amount of an intake valve even when each of the open period and lift amount of the first exhaust valve takes a maximum level obtainable within a control range thereof.

[q] The valve control apparatus as described in the above item [a], wherein the connection changeover mechanism (36) is configured to connect and disconnect the first swing arm (30) with/from the second swing arm (31) when base circular portions of the swing cam (7) and the second drive cam (13,50) are causing the first engine valve (3a) and the second engine valve (3b) to be in a closed state.

That is, motions of both the swing arms (30,31) are in a stopped state when both of the first engine valve (3a) and the second engine valve (3b) are in the closed state. At this time, the connection changeover mechanism (36) can stably connects and disconnects the first swing arm (30) with/from the second swing arm (31).

[r] The valve control apparatus as described in the above item [a], wherein a lift amount of the first engine valve (3a) is controlled to become small at the time of low rotational speed of the engine and to become large at the time of high rotational speed of the engine.

[s] The valve control apparatus as described in the above item [a], wherein the connection changeover mechanism (36) is configured to connect and disconnect the first swing arm (30) with/from the second swing arm (31) in accordance with a rotational speed of the engine.

Accordingly, the output power can be adjusted by switching between the connected state and the unconnected state of the connection changeover mechanism (36) in accordance with the rotational speed of the engine.

This application is based on prior Japanese Patent Application No. 2012-201121 filed on Sep. 13, 2012. The entire contents of this Japanese Patent Application are hereby incorporated by reference.

The scope of the present invention is defined with reference to the following claims.