Abstract:
Disclosed is a control device comprising an IVWAV or EVWAV mechanism. The IVWAV mechanism varies a working angle of an intake valve and the EVWAV varies a working angle of an exhaust valve. An IVOPV mechanism varies an operation phase of the intake valve. An EVOPV mechanism varies an operation phase of the exhaust valve, and a control unit controls the IVWAV or EVWAV mechanism and the IVOPV and EVOPV mechanisms according to an operating condition of an engine. The control unit is configured to control the IVWAV or EVWAV mechanism and the IVOPV and EVOPV mechanisms.

Description:
BACKGROUND OF INVENTION 
     1. Field of Invention 
     The present invention relates in general to a control device for controlling an internal combustion engine, and more particularly to a variable valve control device of an internal combustion engines, which comprises a working angle varying mechanism for varying a working angle of the intake or exhaust valve and an operation phase varying mechanism for varying an operation phase of the intake or exhaust valve. 
     2. Description of Related Art 
     Hitherto, various types of variable valve control devices have been proposed and put into practical use in the field of automotive internal combustion engines. One of such devices is shown in an instruction manual of Toyota car (ALTEZZA) issued on October, 1998 from Toyata Jidosha Kabushiki Kaisha, which comprises generally a so-called intake valve operation phase varying mechanism which varies the operation phase of each intake valve by changing a relative angular position between an intake valve cam shaft and a cam pulley synchronously rotated with the engine crankshaft, and a so-called exhaust valve operation phase varying mechanism which varies the operation phase of each exhaust valve by changing a relative angular position between an exhaust valve cam shaft and the above-mentioned cam pulley. The intake and exhaust valve operation phase varying mechanisms are both powered commonly by a hydraulic pressure produced by an oil pump driven by the engine crankshaft. 
     It is now to be noted that the term “operation phase” used in the description corresponds to the operation timing of the corresponding intake or exhaust valve with respect to that of the engine crankshaft, and the term “working angle” used in the description corresponds to the open period of the corresponding intake or exhaust valve and is represented by an angle range (viz., crank angle) of the engine crankshaft. 
     SUMMARY OF THE INVENTION 
     In general, when, in a middle-load operation range of the engine, a certain valve overlap is provided at or near the top dead center (TDC) on the intake stroke, a certain amount of internal EGR is obtained, which induces reduction in pumping loss and improvement in fuel consumption and exhaust performance. Furthermore, when, in the middle-load operation range, a certain minus valve overlap is provided, a certain amount of exhaust gas is confined in the combustion chamber, which induces reduction in pumping loss and improvement in fuel consumption. It is to be noted that the valve overlap is a phenomenon wherein both the intake and exhaust valves show their open condition simultaneously for a certain time, and the minus valve overlap is a phenomenon wherein both the intake and exhaust valves show their closed condition simultaneously for a certain time. 
     While, in a very low load operation range, such as in the operation range at the time of engine idling, it is necessary to remove or at least minimize the valve overlap and/or minus valve overlap in order to suppress unstable combustion caused by the residual gas of the internal EGR. Accordingly, in case of shifting from the middle-load operation range to the very low-load operation range, such as, in case of rapid deceleration of the engine speed, speedy reduction or cancellation of the valve overlap or minus valve overlap is needed. 
     Accordingly, an object of the present invention is to provide an intake valve control device of an internal combustion engine, which comprises operation phase varying mechanisms for varying an operation phase of the intake and exhaust valves respectively and a working angle varying mechanism for varying a working angle of the intake or exhaust valve, so that in case of engine operation change from a middle-load operation range to a very low-load operation range, reduction or cancellation of the valve overlap and/or minus valve overlap is assuredly and speedily carried out. 
     In order to embody the present invention, the following facts have been seriously considered by the applicants. 
     In a working angle varying mechanism, the biasing force of each valve spring affects to operation of the mechanism. That is, the opening action of the valve is carried out against the biasing force of the valve spring and the closing action of the valve is carried out with the aid of the biasing force. This means that in case of reducing the working angle of the valve, the work of the mechanism is assisted by the biasing force of the valve spring. Thus, under the same hydraulic power applied to the mechanism, responsiveness in such working angle reducing case is higher than that in case of increasing the working angle. 
     While, in an operation phase varying mechanism, a torque is applied to a drive shaft or cam shaft which drives the valve to open and close the same. This means that in case of retarding the operation phase, the work of the mechanism is assisted by the torque. Thus, under the same hydraulic power applied to the mechanism, responsiveness in such operation phase retarding case is higher than that in case of advancing the operation phase. 
     That is, the degree of the responsiveness is represented by the following order. 
     Slow: Increasing a working angle by using the working angle varying mechanism. 
     Slightly fast: Advancing an operation phase by using the operation phase varying mechanism. 
     Fast: Retarding an operation phase by using the operation phase varying mechanism. 
     Very fast: Reducing a working angle by using the working angle varying mechanism. 
     Taking these facts into consideration, the present invention provides a variable valve control device of an internal combustion engine, which, in case of the shifting from the middle-load operation range to the very low-load operation range, selectively operates the operation phase and working angle varying mechanisms in a manner to effectively and speedily reduce or cancel the valve overlap or minus valve overlap. 
     According to a first aspect of the present invention, there is provided a variable valve control device of an internal combustion engine having intake and exhaust valves. The control device comprises an IVWAV mechanism which varies a working angle of the intake valve; an IVOPV mechanism which varies an operation phase of the intake valve; an EVOPV mechanism which varies an operation phase of the exhaust valve; and a control unit which controls the IVWAV, IVOPV and EVOPV mechanisms in accordance with an operation condition of the engine, the control unit being configured to carry out controlling, in a middle-load operation range of the engine, the IVWAV, IVOPV and EVOPV mechanisms to achieve a valve overlap wherein near the top dead center (TDC) on the intake stroke, there is a certain period when both the intake and exhaust valves assume their open conditions, and in case of shifting of the engine from the middle-load operation range to a very low-load operation range, controlling the IVWAV mechanism to reduce the working angle of the intake valve thereby to retard the open timing of the intake valve and controlling the EVOPV mechanism to advance the operation phase of the exhaust valve thereby to advance the close timing of the exhaust valve. 
     According to a second aspect of the present invention, there is provided a variable valve control device of an internal combustion engine having intake and exhaust valves. The control device comprises an IVWAV mechanism which varies a working angle of the intake valve; an IVOPV mechanism which varies an operation phase of the intake valve; an EVOPV mechanism which varies an operation phase of the exhaust valve; and a control unit which controls the IVWAV, IVOPV and EVOPV mechanisms in accordance with an operation condition of the engine, the control unit being configured to carry out controlling, in a middle-load operation range of the engine, the IVWAV, IVOPV and EVOPV mechanisms to achieve a minus valve overlap wherein near the top dead center on the intake stroke, there is a certain period when both the intake and exhaust valves assume their close conditions; and in case of shifting of the engine from the middle-load operation range to a very low-load operation range, controlling the IVOPV mechanism to advance the operation phase of the intake valve thereby to advance the open timing of the intake valve and controlling the EVOPV mechanism to retard the operation phase of the exhaust valve thereby to retard the close timing of the exhaust valve. 
     According to a third aspect of the present invention, there is provided a variable valve control device of an internal combustion engine having intake and exhaust valves. The control device comprises an IVOPV mechanism which varies an operation phase of the intake valve; an EVWAV mechanism which varies a working angle of the exhaust valve; an EVOPV mechanism which varies an operation phase of the exhaust valve; a control unit which controls the IVOPV, EVWAV and EVOPV mechanisms in accordance with an operation condition of the engine, the control unit being configured to carry out controlling, in a middle-load operation range of the engine, the IVOPV, EVWAV and EVOPV mechanisms to achieve a minus valve overlap wherein near the top dead center on the intake stroke, there is a certain period when both the intake and exhaust valves assume their close conditions; and in case of shifting of the engine from the middle-load operation range to a very low-load operation range, controlling the IVOPV mechanism to advance the operation phase of the intake valve thereby to advance the open timing of the intake valve and controlling the EVOPV mechanism to retard the operation phase of the exhaust valve thereby to retard the close timing of the exhaust valve. 
     According to a fourth embodiment of the present invention, there is provided a variable valve control device of an internal combustion engine having intake and exhaust valves. The control device comprises at least one of IVWAV and EVWAV mechanisms, the IVWAV mechanism functioning to vary a working angle of the intake valve and the EVWAV mechanism functioning to vary a working angle of the exhaust valve; an IVOPV mechanism which varies an operation phase of the intake valve; an EVOPV mechanism which varies an operation phase of the exhaust valve; and a control unit which controls the selected one of the IVWAV and EVWAV mechanisms and the IVOPV and EVOPV mechanisms in accordance with an operation condition of the engine, the control unit being configured to carry out controlling, in a middle-loaded operation range of the engine, the selected one of the IVWAV and EVWAV mechanisms and the IVOPV and EVOPV mechanisms to achieve a valve overlap or a minus valve overlap near the top dead center (TDC) on the intake stroke, and in case of shifting of the engine from the middle-load operation range to a very low-load operation range, controlling the IVWAV mechanism or the IVOPV mechanism to shift the open timing of the intake valve toward the top dead center (TDC) on the intake stroke, and controlling the EVWAV mechanism or EVOPV mechanism to shift the close timing of the exhaust valve toward the top dead center (TDC) on the intake stroke. 
    
    
     BRIEF DESCRIPTION OF DRAWINGS 
     FIG. 1 is a perspective view of a variable valve control device of an internal combustion engine, which embodies the present invention; 
     FIG. 2 is a sectional view of the variable valve control device of the invention, showing a part where a working angle varying mechanism is arranged; 
     FIG. 3 is a schematic view of the working angle varying mechanism, which is taken from the direction of an arrow “III” of FIG. 1; 
     FIG. 4 is a diagram showing a hydraulic actuator and a solenoid valve which are used for controlling a control shaft of the working angle varying mechanism; 
     FIG. 5 is an exploded view of an operation phase varying mechanism employed in the variable valve control device of the invention; 
     FIG. 6 is a sectional view of the operation phase varying mechanism in an assembled condition; 
     FIG. 7 is a sectional view of an essential portion of the operation phase varying mechanism; 
     FIG. 8 is a partial view showing an unlocked condition of the operation phase varying mechanism; 
     FIG. 9 is a view similar to FIG. 8, but showing a locked condition of the operation phase varying device; 
     FIGS. 10A and 10B are illustrations showing different conditions of the engine, which are achieved by a first embodiment of the variable valve control device of the invention; 
     FIGS. 11A and 11B are illustrations similar to FIGS. 10A and 10B, but showing the conditions of the engine, which are achieved by a second embodiment of the invention; 
     FIGS. 12A and 12B are illustrations similar to FIGS. 10A and 10B, but showing the conditions of the engine, which are achieved by a third embodiment of the present invention; and 
     FIGS. 13A and 13B are illustrations similar to FIGS. 10A and 10B, but showing the conditions of the engine, which are achieved by a fourth embodiment of the present invention. 
    
    
     DETAILED DESCRIPTION OF EMBODIMENTS 
     In the following, a variable valve control device of the present invention will be described in detail with reference to the accompanying drawings. For ease of understanding, various directional terms such as, right, left, upper, lower, rightward, etc., are used in the description. However, such terms are to be understood with respect to only a drawing or drawings on which the corresponding element or part is illustrated. 
     As will become apparent as the description proceeds, the variable valve control device of the invention is so explained as to be applied to an internal combustion engine having cylinders each having two intake valves and two exhaust valves. For simplification of explanation, the following description is made with respect to only a part of the variable valve control device, which is associated with one of the cylinders of the engine. 
     Referring to FIGS. 1 to  3 , particularly FIG. 1, there is shown one unit (which will be referred to “internal valve control device” hereinafter) of the variable valve control device of an internal combustion engine, which is applied to the intake valves of the engine. 
     It is to be noted that substantially same unit (which will be referred to “exhaust valve control device” hereinafter) is provided by the control device, which is applied to the exhaust valves of the engine. 
     As is seen from FIG. 1, the intake valve control device generally comprises a working angle varying mechanism  1  which varies a working angle of a pair of intake valves  12  of each cylinder, and an operation phase varying mechanism  2  which varies the operation phase of intake valves  12 . 
     As will described in detail in the following, in the working angle varying mechanism  1 , there is arranged a link mechanism by which a drive shaft  13  driven by a crankshaft (not shown) of an associated internal combustion engine through operation phase varying mechanism  2  and two swing cams  20  actuating valve lifters  19  of intake valves  12  to make open/close movement of intake valves  12  against valve springs (not shown) are mechanically linked to continuously vary the working angle (and the valve lift degree) of intake valves  12  while keeping the center point of the working angle constant. It is to be noted that drive shaft  13  extends in a direction along which the cylinders of the engine are aligned. 
     That is, the working angle varying mechanism  1  comprises an eccentric cam  15  eccentrically fixed to drive shaft  13 , a ring-like link  25  rotatably disposed on eccentric cam  15 , a control shaft  16  extending in parallel with drive shaft  13 , a control cam  17  eccentrically fixed to control shaft  16 , a rocker arm  18  rotatably disposed on control cam  17  and having one end  18   b  (see FIG. 2) pivotally connected through a connecting pin  21  to a leading end  25   b  of ring-like link  25 , and a rod-like link  26  by which the other end  18   c  of rocker arm  18  and one of swing cams  20  are linked. 
     As is seen from FIG. 2, the center “X” of eccentric cam  15  is displaced from the center “Y” of drive shaft  13  by a predetermined degree, and the center “P 1 ” of control cam  17  is displaced from the center “P 2 ” of control shaft  16  by a predetermined degree. As is seen from FIGS. 2 and 3, a journal portion  20   b  of swing cam  20 , which is rotatably disposed about drive shaft  13 , and a journal portion of control shaft  16  are rotatably held by a pair of brackets  14   a  and  14   b  which are secured to a cylinder head  11  of the engine through common bolts  14   c.    
     As is seen from FIG. 1, the rod-like link  26  is arranged to extend generally along an axis of the corresponding intake valve  12 . As is seen from FIG. 2, one end  26   a  of rod-like link  26  is pivotally connected to the other end  18   c  of rocker arm  18  through a connecting pin  28 . 
     When, with the above-mentioned arrangement, the drive shaft  13  is rotated due to rotation of crankshaft, the ring-like link  25  is forced to make a translation motion through eccentric cam  15 , and thus the swing cam  20  is forced to swing through rocker arm  18  and rod-like link  26  resulting in that the intake valves  12  are forced to make open/close movement against force of the valve springs (not shown). 
     While, when the control shaft  16  is rotated within a given angular range by an after-mentioned actuator  30 , the center “P 1 ” of control cam  17 , which serves as a rotation center of rocker arm  18 , is forced to move about the center “P 2 ” of control shaft  16 . With this movement, a link unit including ring-like link  25 , rocker arm  18  and rod-like link  26  is forced to change its posture and thus the working angle and valve lift degree of intake valves  12  are continuously varied keeping the operation phase of the same constant. 
     In the above-mentioned working angle varying mechanism  1 , the swing cam  20  which actuates intake valve  12  is rotatably disposed about drive shaft  13  which is rotated along with the crankshaft of the engine. Accordingly, undesired center displacement of swing cam  20  relative to drive shaft  13  is suppressed, and thus, controllability is improved. Since the swing cam  20  is supported by drive shaft  13 , there is no need of providing a separate supporting shaft for swing cam  20 . Thus, advantages are expected in view of the number of parts used and the mounting space. Furthermore, since the connecting portions of the parts are made through a so-called surface to surface contact, adequate abrasion resistance is obtained. 
     Referring to FIG. 4, there is shown the actuator  30  which rotates control shaft  16  within a predetermined angular range. The actuator  30  comprises a cylinder  39  of which interior is divided into first and second hydraulic chambers  33  and  34  due to provision of a piston proper part  32   a  of a piston  32 . Thus, in accordance with a pressure difference appearing between the first and second hydraulic chambers  33  and  34 , the piston  32  is forced to move in a fore-and-aft direction. A stem portion of piston  32  has a leading end exposed to the open air. The leading end of the piston stem has a pin  32   b  fixed thereto. As shown, the piston stem extends perpendicular to an axis of control shaft  16 . A link plate  16   a  is fixed to one end of control shaft  16  to rotate therewith about the axis of control shaft  16 . The link plate  16   a  is formed with a radially extending slot  16   b  with which the pin  32   b  of the piston stem is slidably engaged. Accordingly, upon the fore-and-aft movement of piston  32 , the control shaft  16  is rotated within a predetermined angular range about its axis. 
     Oil supply to first and second hydraulic chambers  33  and  34  is switched in accordance with the position of a spool  35  of a solenoid valve  31 . The solenoid valve  31  is controlled in ON/OFF manner (viz., duty-control) by a control signal issued from an engine control unit  3 . The control unit  3  comprises a microcomputer including generally CPU, RAM, ROM and input and output interfaces. That is, by varying the duty ratio of the control signal in accordance with the operation condition of the engine, the position of spool  35  is changed. 
     That is, when, as shown in the drawing, the spool  35  assumes a rightmost position, a first hydraulic passage  36  connected with first hydraulic chamber  33  is connected with an oil pump  9  thereby feeding first hydraulic chamber  33  with a hydraulic pressure and at the same time, a second hydraulic passage  37  connected with second hydraulic chamber  34  is connected with a drain passage  38  thereby draining the oil from second hydraulic chamber  34 . Accordingly, the piston  32  of actuator  30  is shifted leftward in the drawing. 
     While, when the spool  35  assumes a leftmost position in the drawing, the first hydraulic passage  36  is connected with drain passage  38  to drain the oil from first hydraulic chamber  33 , and at the same time, the second hydraulic passage  37  is connected with oil pump  9  to feed second hydraulic chamber  34  with a hydraulic pressure. Thus, the piston  32  is shifted rightward in the drawing. 
     While, when the spool  35  is in a middle position, both of first and second hydraulic passages  36  and  37  are closed by spool  35 , and thus, the hydraulic pressure in first and second hydraulic chambers  33  and  34  is held or locked thereby holding piston  32  in a corresponding middle position. 
     As is described hereinabove, the piston  32  of actuator  30  is moved to or held at a desired position, and thus, the working angle of intake valves  12  can be controlled to a desired angle within a predetermined angular range. 
     It is to be noted that the engine control unit  3  controls working angle varying mechanism  1  and operation phase varying mechanism  2  in accordance with an engine speed, an engine load, a temperature of engine cooling water and a vehicle speed. In addition to this control, the engine control unit  3  carries out an ignition timing control, a fuel supply control, a transition correction control and a fail-safe control. 
     In the following, the operation phase varying mechanism  2  will be described with reference to FIGS. 5 to  9  and FIG.  1 . 
     As will become apparent as the description proceeds, the operation phase varying mechanism  2  functions to vary a relative angular position between drive shaft  13  and a timing pulley  40  that is rotatably disposed on drive shaft  13  and synchronously rotated together with the engine crankshaft, so that the operation phase of intake valves  12  is varied while keeping the working angle and the valve lift degree of intake valves  12  constant. 
     That is, as is seen from FIGS. 1,  5  and  6 , the operation phase varying mechanism  2  comprises generally the timing pulley  40  fixed to an axial end of drive shaft  13 , a vane unit  41  rotatably installed in timing pulley  40  and a hydraulic circuit structure arranged to rotate vane unit  41  in both directions by a hydraulic power. 
     As is seen from FIG. 5, the timing pulley  40  generally. comprises a rotor member  42  which has an external gear  42   a  meshed with teeth of a timing chain (not shown), a cylindrical housing  43  which is arranged in front of rotor member  42  and rotatably disposes therein vane unit  41 , a circular front cover  44  which covers a front open end of the housing  43 , a circular rear cover  45  which is arranged between housing  43  and rotor member  42  and covers a rear open end of housing  43 , and a plurality of bolts  46  (see FIG. 6) which coaxially connects housing  43 , front cover  44  and rear cover  45  as a unit. 
     As is seen from FIGS. 5 and 6, the rotor member  42  is of a cylindrical member and has a center bore  42   a  formed therethrough. The rotor member  42  is formed with a plurality of internally threaded bolt holes (no numerals) with which the threads of bolts  46  are engaged. Furthermore, as is seen from FIG. 6, the center bore  42   a  of rotor member  42  has a diametrically enlarged rear (or right) portion  48  which is mated with an after-mentioned sleeve member  47 . Furthermore, the rotor member  42  has at its front (or left) side a coaxial circular recess  49  which has rear cover  45  mated therewith. The rotor member  42  has further an engaging hole  50  at a given portion of circular recess  49 . 
     As is seen from FIG. 5, the cylindrical housing  43  has axial both ends opened and has on its inner surface four axially extending partition ridges  51  which are arranged at equally spaced intervals (viz., 90°). As shown, each partition ridge  51  has a generally trapezoidal cross section and has axial both ends flush with the both ends of cylindrical housing  43 . Furthermore, each partition ridge  51  has an axially extending bolt hole  52  through which the corresponding bolt  46  passes. Furthermore, each partition ridge  51  has at its inner top portion an axially extending holding groove  51   a . As may be seen from FIG. 6, each holding groove  51   a  receives therein an elongate seal member  53  and a plate spring  54  which biases seal member  53  radially inwardly. 
     As is seen from FIG. 5, the circular front cover  44  is formed with a center opening  55 . The front cover  44  further has four bolt holes (no numerals) which are mated with bolt holes  52  of the cylindrical housing  43 . 
     As is seen from FIG. 5, the circular rear cover  45  is formed on its rear side with an annular ridge  56  which is intimately engaged with circular recess  49  of the above-mentioned rotor member  42 . Furthermore, the rear cover  45  is formed with a center opening  57  with which a smaller diameter annular portion  56  of sleeve member  47  is engaged. The rear cover  45  has further four bolt holes (no numerals) which are mated with bolt holes  52  of cylindrical housing  43 . Furthermore, the rear cover  45  is formed with an engaging hole  50 ′ at a position corresponding to engaging hole  50  of rotor member  42 . 
     As is seen from FIG. 5, the vane unit  41  is made of a sintered alloy and is connected to the front end of drive shaft  13  (see FIG. 1) through a connecting bolt  58 . That is, the vane unit  41  is rotated together with drive shaft  13 . More specifically, the vane unit  41  comprises a cylindrical base portion  59  which has an axially extending bore  41   a  through which the connecting bolt  58  passes, and four equally spaced and axially extending vane portions  60  which are raised radially outward from base portion  59 . 
     As shown, each vane portion  60  is in the rectangular shape, and as is seen from FIG. 7, each vane portion  60  is put between two adjacent partition ridges  51  of housing  43 . Each vane portion  60  has at its outer top portion an axially extending holding groove  61 . Each holding groove  61  receives therein an elongate seal member  62  and a plate spring  63  which biases seal member  62  radially outwardly. As shown in FIG. 7, each seal member  53  of cylindrical housing  43  is biased against an outer cylindrical wall of the cylindrical base portion of vane unit  41  to establish a hermetic sealing therebetween, and each seal member  62  of vane unit  41  is biases against an inner cylindrical wall of cylindrical housing  43  to establish a hermetic sealing therebetween. 
     As is seen from FIG. 7, due to placement of the vane portion  60  of vane unit  41  in each space defined between two adjacent partition ridges  51  of cylindrical housing  43 , there are defined an advancing hydraulic chamber  64  and a retarding hydraulic chamber  65  in the space. 
     As is seen from FIGS. 5 and 7, one of vane portions  60  of the vane unit  41  is formed with an axially extending bore  66  at a position corresponding to the engaging hole  50 ′ of rear cover  45 . As is seen from FIG. 5, the vane portion  60  is formed with a small passage  67  for connecting advancing and retarding hydraulic chambers  65  and  66 . 
     As is seen from FIGS. 5 and 6, a lock pin  68  is axially slidably received in the axially extending bore  66  of vane portion  60 . As is seen from FIGS. 8 and 9, the lock pin  68  comprises a cylindrical middle portion  68   a , a smaller diameter engaging portion  68   b  and a larger diameter stopper portion  68   c.    
     As is seen from FIG. 8, for hydraulically actuating lock pin  68  in bore  66  of vane portion  60 , there is formed a pressure receiving chamber  69  which is defined by a stepped surface of the larger diameter stopper portion  68   c , the an outer surface of middle portion  68   a  and a cylindrical inner wall of bore  66 . Between the lock pin  68  and the front cover  44 , there is compressed a coil spring  70  which biases the lock pin  68  toward the rear cover  45 . 
     It is to be noted that when the vane unit  41  assumes a most retarded angular position, the engaging portion  68   b  of the lock pin  68  is engaged with the engaging hole  50 ′ of the rear cover  45  as is seen from FIG.  9 . 
     As is seen from FIG. 6, the hydraulic circuit structure comprises a first hydraulic passage  71  through which hydraulic pressure is fed to or discharged from the advancing hydraulic chamber  64  and a second hydraulic passage  72  through which hydraulic pressure is fed to or discharged from the retarding hydraulic chamber  65 . These first and second hydraulic passages  71  and  72  are connected to supply and drain passages  73  and  74  through an electromagnetic switch valve  75 . 
     As is seen from FIG. 6, the first hydraulic passage  71  comprises a first passage part  71   a  which is formed in both cylinder head  11  and drive shaft  13 , a first oil passage  71   b  which is formed in the connecting bolt  58  and connected to first passage part  71   a , an oil chamber  71   c  which is defined between an outer cylindrical surface of an enlarged head of the connecting bolt  58  and an inner cylindrical surface of the axially extending bore  41   a  of base portion  59  of vane unit  41  and connected to first oil passage  71   b  and four radially extending branched passages  71   d  which are formed in base portion  59  of vane unit  41  to connect the oil chamber  71   c  with the four advancing hydraulic chambers  64 . 
     While, as is seen from FIG. 6, the second hydraulic passage  72  comprises a second passage part  72   a  which is formed in both cylinder head  11  and drive shaft  13 , a second oil passage  72   b  which is formed in sleeve member  57  and connected to second passage part  72   a , four oil grooves  72   c  formed at an inner surface of center bore  42   a  of rotor member  42  and connected to second oil passage  72   b  and four oil holes  72   d  which are formed in rear cover  45  at equally spaced intervals to connect the four oil grooves  72   c  with four retarding hydraulic chambers  65  respectively. 
     The electromagnetic switch valve  75  is of a type having four ports and three operation positions. That is, due to movement of a spool installed in valve  75 , the first and second hydraulic passages  71  and  72  are selectively connected to and blocked from supply and drain passages  73  and  74 . The movement of the spool is controlled (duty-control) by a control signal issued from engine control unit  3 . 
     By processing information signals from a crank angle sensor and an air flow meter, the control unit  3  detects an existing operation condition of the engine. Furthermore, by processing information signals from a crank angle sensor and a cam angle sensor, the control units  3  detects a relative angular position between timing pulley  40  and drive shaft  13 . 
     In an initial condition induced when the engine stops, the spool of valve  75  assumes its rightmost position as shown in FIG.  6 . In this condition, the supply passage  73  is connected with second hydraulic passage  72  and at the same time, the drain passage  74  is connected with first hydraulic passage  71 . Accordingly, hydraulic pressure in the four retarding hydraulic chambers  65  is kept unchanged, while hydraulic pressure in the four advancing hydraulic chambers  64  is reduced to zero due to connection with drain passage  74 . Under this condition, as is seen from FIG. 7, the vane unit  41  assumes a leftmost position or most retarded position wherein each vane portion  60  abuts against a right face of the corresponding left partition ridge  51  of cylindrical housing  43 . In this condition, the operation phase of each intake valve  12  is controlled at a retarded side. 
     In an initial stage of engine starting, the vane unit  41  is held in the most retarded position. When, under this initial stage, the hydraulic pressure in the retarding hydraulic chambers  65  is relatively low in such a degree that the hydraulic pressure fed to pressure receiving chamber  69  through bore  67  is still lower than the force of coil spring  70 , the lock pin  68  is kept engaged with engaging hole  50 ′ of the rear cover  45 , as is shown in FIG.  9 . Accordingly, the vane unit  41  is locked to cylindrical housing  43  keeping the most retarded angular position. Thus, undesired vibration, which would be caused by a varying hydraulic pressure in the retarding hydraulic chambers  64  and a varying torque produced by drive shaft  13 , is suppressed or at least minimized. This prevents generation of noises caused by collision of vane portions  60  against partition ridges  51 . 
     When, after passing of a certain time from the engine starting, the hydraulic pressure in retarding hydraulic chamber  65  is increased and at the same time the hydraulic pressure in pressure receiving chamber  69  is increased. Thus, the lock pin  68  is moved back against the force of coil spring  70  and thus finally, as is seen from FIG. 8, the lock pin  68  is disengaged from engaging hole  50 ′ of rear cover  45 . Upon this, the locked condition between vane unit  41  and cylindrical housing  43  becomes canceled permitting free rotation of vane unit  41  in the housing  43 . 
     When the spool (see FIG. 6) of the switch valve  75  is moved to its leftmost position in the drawing, the supply passage  73  becomes connected with first hydraulic passage  71  and at the same time the drain passage  74  becomes connected with second hydraulic passage  72 . Accordingly, in this condition, hydraulic pressure in the retarding hydraulic chamber  65  is led to the oil pan through second hydraulic passage  72  and drain passage  74 , and at the same time, hydraulic pressure from the oil pump  9  is led into advancing hydraulic chamber  64  through supply passage  73  and first hydraulic passage  71 . Upon this, the vane unit  41  is turned in a clockwise direction in FIG. 7, that is, in an advancing direction, and thus, the operation phase of each intake valve  12  is shifted to an advanced side. 
     While, when the spool (see FIG. 6) of switch valve  75  is kept in a middle position, both first and second hydraulic passages  71  and  72  are blocked by the spool. As a result, hydraulic pressure in both first and second hydraulic chambers  33  and  34  of actuator  30  are locked, so that the vane unit  41  assumes a corresponding intermediate position, keeping the operation phase of each intake valve  12  at a corresponding value. 
     As is described hereinabove, in the operation phase varying mechanism  2 , by changing the position of the spool of electromagnetic switch valve  75  in accordance with the operation condition of the engine, the vane unit  41  can be held in a desired intermediate position. That is, according to the operation phase varying mechanism  2 , the operation phase of each intake valve  12  can be varied and held in a desired value irrespective of the simple structure possessed by mechanism  2 . 
     As is easily seen from FIG. 1, in the intake valve control device of the invention, the working angle varying mechanism  1  and the operation phase varying mechanism  2  are arranged at different positions without making a relative interference therebetween. Both the mechanisms  1  and  2  are powered by a common oil pump  9 , which is one of conditions to simplify the construction of the intake valve control device. 
     As has been described hereinabove, the exhaust valve control device has substantially the same construction as the above-mentioned intake valve control device. That is, the above description on the intake valve control device can be equally applied to the exhaust valve control device except the type of the valves. That is, in case of the exhaust valve control device, the valves  12  (see FIG. 1) actuated by the swing cams  20  are a pair of exhaust valves of the associated engine. 
     For ease of understanding, the working angle and operation phase varying mechanisms for the exhaust valves will be denoted by ( 1 ) and ( 2 ) and the exhaust valves actuated by these mechanisms ( 1 ) and ( 2 ) will be denoted by ( 12 ). 
     FIGS. 10A and 10B are illustrations schematically showing open/close timing of the intake and exhaust valves, which is provided by a first embodiment of the present invention. 
     In this first embodiment, controlling of intake valves  12  is carried out by allowing control unit  3  to control both the working angle and operation phase varying mechanisms  1  and  2  for intake valves  12 , and controlling of exhaust valves ( 12 ) is carried out by allowing control unit  3  to control operation phase varying mechanism ( 2 ) for exhaust valves ( 12 ). 
     As shown in FIG. 10A, in a middle-load operation range, the open timing of intake valve  12  is set before the top dead center (TDC) on the intake stroke, and the close timing of exhaust valve ( 12 ) is set after the top dead center (TDC) on the intake stroke, so that in the vicinity of the top dead center (TDC) on the intake stroke, there is produced a valve overlap of a degree “ΔD 1 ”. With this production, a certain amount of internal EGR gas is obtained inducing reduction in pumping loss and improvement in fuel consumption. 
     While in a very low-load operation range, such as a range induced when the engine is under idling, such valve overlap is removed for improving the combustion stability. 
     Accordingly, in case of rapid shifting of the engine from the middle-load operation range to the very low-load operation range, such as, in case of rapid deceleration of the engine speed, speedy reduction or cancellation of the valve overlap is needed. 
     Thus, in the first embodiment, upon need of this speedy reduction of the valve overlap, the open timing of the intake valve  12  is retarded toward the top dead center (TDC) on the intake stroke and at the same time the close timing of the exhaust valve ( 12 ) is advanced toward the top dead center (TDC) on the intake stroke. 
     For retarding the open timing of intake valve  12 , there are two methods, one being a method executed by working angle varying mechanism  1  for intake valves  12 , and the other being a method executed by operation phase varying mechanism  2  for intake valves  12 . In the method by mechanism  1 , the working angle of intake valve  12  is reduced, and in the method by the other mechanism  2 , the operation phase of intake valve  12  is retarded. 
     In case of reducing the working angle of intake valve  12  by working angle varying mechanism  1 , the valve spring for intake valve  12  assists the needed work of mechanism  1 , and thus, satisfied responsiveness in working angle change is obtained by mechanism  1 . Accordingly, upon need of the rapid shifting from the middle-load operation range to the very low-load operation range, the working angle varying mechanism  1  is actuated to reduce the working angle of intake valve  12  while stopping operation of operation phase varying mechanism  2 . With this, the open timing of intake valve  12  is speedily retarded. 
     While, in case of advancing the close timing of exhaust valve ( 12 ), the operation phase varying mechanism ( 2 ) is actuated. In this mechanism ( 2 ), since the cam shaft or the drive shaft ( 13 ) is constantly applied with a certain torque, having the operation phase advanced needs a certain hydraulic pressure that overcomes the torque of drive shaft ( 13 ). Accordingly, upon need of rapid shifting from the middle-load operation range to the very low-load operation range, the hydraulic pressure is instantly fed to operation phase varying mechanism ( 2 ) to instantly and effectively actuate mechanism ( 2 ). With this, the close timing of exhaust valve ( 12 ) is speedily advanced. 
     That is, upon need of the above-mentioned rapid shifting, retardation of the open timing of intake valves  12  is effected by the working angle varying mechanism  1  for intake valves  12 , and at the same time, advancement of the close timing of exhaust valves ( 12 ) is effected by the operation phase varying mechanism ( 2 ). 
     In order to embody such operation, the following measures are employed in the first embodiment, which will be described with reference to FIGS. 4,  6  and  7 . 
     That is, upon need of such rapid shifting, a condition is produced by control unit  3  (see FIGS. 4 and 6) wherein a practical sectional area of a first hydraulic line (see FIGS. 6 and 7) extending from oil pump  9  to the advancing hydraulic chamber  64  of the operation phase varying mechanism ( 2 ) is greater than a practical sectional area of a second hydraulic line (see FIG. 4) extending from oil pump  9  to the first or second hydraulic chamber  33  or  34  of working angle varying mechanism  1 . 
     More specifically, upon need of the rapid shifting, the duty ratio of a control signal fed to the electromagnetic switch valve  75  (see FIG. 6) of operation phase varying mechanism ( 2 ) is controlled to a highest value (for example 100%) that corresponds to the most advancing degree, and the duty ratio of a control signal fed to solenoid valve  31  (see FIG. 4) of working angle varying mechanism  1  is controlled to an intermediate value that is higher than 0%. However, if desired, the first hydraulic line may be constructed to have a flow resistance that is sufficiently smaller than that of the second hydraulic line. 
     FIGS. 11A and 11B are illustrations schematically showing open/close timing of the intake and exhaust valves  12  and ( 12 ), which is provided by a second embodiment of the present invention. 
     Similar to the above-mentioned first embodiment, in this second embodiment, controlling of intake valves  12  is carried out by allowing control unit  3  to control both working angle and operation phase varying mechanisms  1  and  2  for intake valves  12 , and controlling of exhaust valves ( 12 ) is carried out by allowing control unit  3  to control only the operation phase varying mechanism ( 2 ) for exhaust valves ( 12 ). 
     As shown in FIG. 11A, in a middle-load operation range, the open timing of intake valve  12  is set after the top dead center (TDC) on the intake stroke and the close timing of exhaust valve ( 12 ) is set before the top dead center (TDC) on the intake stroke, so that in the vicinity of the top dead center (TDC) on the intake stroke, there is produced a minus valve overlap of a degree “ΔD 2 ”. With this production, a certain amount of exhaust gas is left in the cylinder in the vicinity of the top dead center (TDC) on intake stroke, so that reduction of pumping loss and improvement in fuel consumption are achieved. 
     In case of rapid shifting of the engine from the middle-load operation range to the very low-load operation range, speedy reduction or cancellation of the minus valve overlap is needed in order to assure a stable combustion in the very low-load operation range. That is, if the residual gas is remained in the very low-load operation range, the engine fails to operate stably. 
     Thus, in the second embodiment, upon need of this speedy reduction of the minus valve overlap, the open timing of intake valve  12  is advanced toward the top dead center (TDC) on the intake stroke and at the same time the close timing of exhaust valve ( 12 ) is retarded toward the top dead center (TDC) on the intake stroke. 
     For advancing the open timing of intake valve  12 , there are two methods, one being a method executed by the working angle varying mechanism  1 , and the other being a method executed by the operation phase varying mechanism  2 . In the method by mechanism  1 , the working angle of intake valve  12  is increased and in the method by the other mechanism  2 , the operation phase of intake valve  12  is advanced. 
     In case of increasing the working angle of intake valve  12  by working angle varying mechanism  1 , the valve spring for intake valve  12  works to obstruct the needed work of mechanism  1 . That is, increasing of the working angle needs a certain hydraulic pressure that overcomes the biasing force of the valve spring. Due to this reason, desired responsiveness in increasing the working angle is not expected. 
     While, in case of advancing the operation phase of intake valve  12  by using operation phase varying mechanism  2 , there is a need of a hydraulic pressure that overcomes the torque applied to drive shaft  13 . However, since, in the middle-load operation range, the working angle is relatively small, the torque of drive shaft  13  is accordingly small, and thus, the hydraulic pressure needed for actuating the mechanism  2  to advance the operation phase of intake valve  12  is controlled to a relatively small value. 
     That is, under an even energy, that is, under the even hydraulic pressure produced by the oil pump  9 , the operation phase varying mechanism  2  can exhibit a higher responsiveness in advancing the open timing of intake valve  12  than the working angle varying mechanism  1 . Accordingly, upon need of the rapid shifting from the middle-load operation range to the very low-load operation range, the operation phase varying mechanism  2  is actuated to advance the operation phase of intake valve  12  while stopping operation of the working angle varying mechanism  1 . With this, the open timing of intake valve  12  is speedily advanced. 
     While, in case of retarding the close timing of exhaust valve ( 12 ), the operation phase varying mechanism ( 2 ) for the exhaust valves ( 12 ) is actuated. Since, in this case, a certain torque constantly applied to the exhaust cam shaft functions to assist the needed movement of exhaust valve ( 12 ), the mechanism ( 2 ) exhibits a higher responsiveness in varying (or retarding) the close timing of exhaust valve ( 12 ) than the mechanism  1  in varying (or advancing) the open timing of intake valve  12 . 
     Accordingly, upon need of the rapid shifting, the hydraulic pressure is instantly fed to the operation phase varying mechanism  2  to instantly and effectively actuate the mechanism  2 . With this, advancing of the open timing of intake valve  12  and retarding of the close timing of exhaust valve ( 12 ) are instantly achieved at the same time. 
     That is, like in the case of the above-mentioned first embodiment, upon need of the rapid shifting, the control unit  3  (see FIGS. 4 and 6) operates to establish a condition wherein the practical sectional area of the first hydraulic line (see FIGS. 6 and 7) extending from oil pump  9  to advancing hydraulic chamber ( 64 ) of operation phase varying mechanism ( 2 ) for exhaust valves ( 12 ) is greater than the practical sectional area of second hydraulic line (see FIG. 4) extending from oil pump  9  to first or second hydraulic chamber  33  or  34  of working angle varying mechanism  1  for intake valves  12 . 
     More specifically, upon need of the rapid shifting, the duty ratio of the control signal fed from control unit  3  to solenoid valve  31  (see FIG. 4) and that of the control signal fed from control unit  3  to electromagnetic switch valve  75  (see FIG. 6) are so controlled as to established the above-mentioned condition. 
     Usually, in the middle-load operation range, the working angle of intake valve  12  is set smaller than that of exhaust valve ( 12 ). Thus, under shifting from the middle-load operation range to the very low-load operation range, the hydraulic power needed by operation phase varying mechanism  2  is controlled relatively small, so that the reduction of the minus valve overlap is effectively made. 
     FIGS. 12A and 12B are illustrations schematically showing open/close timing of the intake and exhaust valves  12  and ( 12 ), which is provided by a third embodiment of the present invention. 
     In this third embodiment, controlling of intake valves  12  is carried out by allowing control unit  3  to control operation phase varying mechanism  2  for intake valves  12 , and controlling of exhaust valves ( 12 ) is carried out by allowing control unit  3  to control both working angle and operation phase varying mechanisms ( 1 ) and ( 2 ) for exhaust valves ( 12 ). 
     As is seen from FIG. 12A, in a middle-load operation range, the open timing of intake valve  12  is set after the top dead center (TDC) on the intake stroke and the close timing of exhaust valve ( 12 ) is set before the top dead center (TDC) on the intake stroke, so that in the vanity of the top dead center (TDC) on the intake stroke, there is produced a minus valve overlap of a degree “ΔD 2 ”. Thus, reduction of pumping loss and improvement in fuel consumption in such middle-load operation range are achieved. 
     Generally, in the middle-load operation range, the working angle of exhaust valve ( 12 ) is set relatively large in order to advance the open timing of exhaust valve ( 12 ) toward the bottom dead center (BDC). 
     Like in the above-mentioned second embodiment, upon need of shifting from the middle-loaded operation range to the very low-load operation range, the open timing of intake valve  12  is advanced toward the top dead center (TDC) on the intake stroke and at the same time the close timing of exhaust valve ( 12 ) is retarded toward the top dead center (TDC) on the intake stroke to speedily reduce or cancel the minus valve overlap. 
     For retarding the close timing of exhaust valve ( 12 ), there are two methods, one being a method executed by working angle varying mechanism ( 1 ), and the other being a method executed by operation phase varying mechanism ( 2 ). In the method by working angle varying mechanism ( 1 ), the working angle of exhaust valve ( 12 ) is increased and in the method by the other mechanism ( 2 ), the operation phase of exhaust valve ( 12 ) is retarded. 
     For the same reason as mentioned in the second embodiment, under an even energy, that is, under the even hydraulic pressure produced by oil pump  9 , the operation phase varying mechanism ( 2 ) can exhibit a higher responsiveness in retarding the close timing of exhaust valve ( 12 ) than working angle varying mechanism ( 1 ). Accordingly, upon need of the rapid shifting from the middle-loaded operation range to the very low-load operation range, the operation phase varying mechanism  2  is actuated to advance the operation phase of intake valve  12  and at the same time the operation phase varying mechanism ( 2 ) is actuated to retard the operation phase of exhaust valve ( 12 ). Since the certain torque constantly applied to the exhaust cam shaft functions to assist the needed movement of exhaust valve ( 12 ), the mechanism ( 2 ) exhibits a higher responsiveness in varying (or retarding) the close timing of exhaust valve ( 12 ) than the mechanism  1  in varying (or advancing) the open timing of intake valve  12 . 
     Accordingly, upon need of the rapid shifting, the hydraulic pressure is instantly fed to the operation phase varying mechanism  2  to instantly and effectively actuate the mechanism  2 . With this, advancing of the open timing of intake valve  12  and retarding of the close timing of exhaust valve ( 12 ) are instantly achieved at the same time. 
     Like in the above-mentioned first and second embodiments, upon need of the rapid shifting, the control unit  3  (see FIGS. 4 and 6) operates to establish a condition wherein the practical sectional area of a first hydraulic line (see FIGS. 6 and 7) extending from oil pump  9  to advancing hydraulic chamber  64  of operation phase varying mechanism  2  for intake valves  12  is greater than the practical sectional area of a second hydraulic line (see FIG. 4) extending from oil pump  9  to retarding hydraulic chamber ( 65 ) of operation phase varying mechanism ( 2 ) for exhaust valves ( 12 ). 
     More specifically, upon rapid shifting from the middle-load operation range to the very low-load operation range, that is, upon a rapid deceleration of the engine, the intake air is reduced due to reduction in engine speed, which induces retardation of the opening timing of exhaust valve ( 12 ) due to a so-called exhaust inertial effect. As is described hereinabove, in the third embodiment, for reducing or canceling the minus valve overlap, the operation phase of exhaust valve ( 12 ) is retarded by operation phase varying mechanism ( 2 ) and at the same time the open timing of the of exhaust valve ( 12 ) is retarded toward the bottom dead center (BDC). That is, in the third embodiment, upon the rapid shifting, there is no need of actuating working angle varying mechanism ( 1 ) for exhaust valves ( 12 ), and thus, energy is saved. 
     FIGS. 13A and 13B are illustrations schematically showing open/close timing of intake and exhaust valves  12  and ( 12 ), which is provided by a fourth embodiment of the present invention. The fourth embodiment is basically the same as the above-mentioned third embodiment except for the following. 
     That is, as is easily understood when comparing FIG.  13 A and FIG. 12A, in the fourth embodiment, in the middle-load operation range, the working angle of exhaust valve ( 12 ) is set smaller than that in the case of the third embodiment and the open timing of exhaust valve ( 12 ) is set near or slightly after the bottom dead center (BDC). 
     Upon need of shifting from the middle-load operation range to the very low-load operation range due to rapid reduction of the engine speed, the operation phase of intake valve  12  is advanced by operation phase varying mechanism  2  for intake valves  12  and at the same time the operation phase of exhaust valve ( 12 ) is retarded by operation phase varying mechanism ( 2 ) for exhaust valves ( 12 ) without varying the working angle of exhaust valve ( 12 ) by the working angle varying mechanism ( 1 ) for exhaust valves ( 12 ). This is similar to the work in the third embodiment. 
     Thus, in the fourth embodiment, upon need of the rapid shifting from the middle-load operation range to the very low-load operation range, the minus valve overlap is effectively and speedily reduced or cancelled, like in the case of the third embodiment. Furthermore, since the open timing of exhaust valve ( 12 ) is retarded in compliance with retardation of the close timing of exhaust valve ( 12 ), a certain engine braking is effectively achieved upon reduction of the engine speed. 
     The entire contents of Japanese Patent Application 2000-262109 (filed Aug. 31, 2000) are incorporated herein by reference. 
     Although the invention has been described above with reference to the embodiments of the invention, the invention is not limited to such embodiments as described above. Various modifications and variations of such embodiments may be carried out by those skilled in the art, in light of the above descriptions.