Abstract:
A system for varying actuation of a combustion valve in an internal combustion engine, including a control shaft pivot housing fixedly disposed on the engine, a control shaft pivotably disposed within the control shaft pivot housing and eccentrically fixed to a control shaft disc, an input rocker subassembly pivotably disposed on the control shaft disc and having a contact feature disposable as a follower against a camshaft lobe, an output cam subassembly pivotably disposed on the control shaft and engageable by the input rocker subassembly and including an output cam profile for engaging a finger follower of the engine, and a bias spring to urge the output cam subassembly toward the input rocker subassembly. In one aspect of the invention, high lift events with full duration are produced whenever the control shaft is aligned such that the input rocker pivot center and the output cam pivot center are coincidental.

Description:
RELATIONSHIP TO OTHER APPLICATIONS AND PATENTS 
     This application claims the benefit from U.S. Provisional Application Ser. No. 61/242,211, filed on Sep. 14, 2009. 
    
    
     TECHNICAL FIELD 
     The present invention relates to Variable Valvetrain Actuation (VVA) devices for varying the lift of combustion valves in an internal combustion engine; more particularly, to such devices for varying the lift, duration, and phasing of such valves; and most particularly, to such devices employing a single rotary actuator, referred to herein as a High Efficiency Lift Profiler (HELP) system. 
     BACKGROUND OF THE INVENTION 
     When compared to competitive VVA devices, the main advantages of the present HELP system are its simplicity and compact height. The VVA device disclosed in US Patent Application Publication No. 2003/0132813 A1 and U.S. Pat. No. 7,246,578 B2 are kinematically complex, adding four to six oscillating members per cylinder to conventional, direct acting and roller finger follower valvetrains, respectively. The greater the number of oscillating parts, the less stiff the system is dynamically and the less likely it is to obtain satisfactory high speed operation. This can be seen in these VVA devices&#39; undesirable phase change characteristic in full lift as a function of engine speed. 
     Although the VVA device disclosed in U.S. Pat. No. 6,823,826 B1 offers an attractive packaging height, it is very complex as well. Moreover, with its internally and externally splined parts, it is a costly and noisy solution for varying valve lift. 
     In the present HELP system, only two oscillating members have been added to a standard, roller finger follower type valvetrain system to effect variations in lift, timing, and duration. By comparison, while the VVA device disclosed in U.S. Pat. No. 7,225,773 B2 also is kinematically simple with only two added parts per cylinder, the system adds considerable height to an engine. 
     Having fewer moving parts also simplifies the design tradeoffs associated with these mechanical devices. In the VVA device disclosed in U.S. Pat. No. 5,937,809, an input rocker is connected through a link to two output cams that also ride on the input camshaft. Because the mechanism comprises four moving parts per cylinder, it is difficult to provide a return spring stiff enough for high-speed engine operation that will still fit within the available packaging space. Since the present HELP system has only two moving parts, the total mass moment of inertia is much lower, and hence spring design is less challenging. In addition, because there are fewer parts, there are fewer degrees of freedom in the mechanism, which simplifies the task of design optimization to meet performance criteria by substantially reducing the number of equations required to describe the mechanism&#39;s motion. 
     As the cost of petroleum continues to fluctuate from increased global demands and limited supplies, the fuel economy benefits of internal combustion engines will become a central issue in their design, manufacture, and use at the consumer level. Production applications that apply a continuously variable valvetrain system to just the intake side of a gasoline engine can yield notable fuel economy benefits on FTP (Federal Test Procedure—USA) or NEDC (New European Driving Cycle) driving schedules, based on simulations and real vehicle testing. VVA when combined with Direct Injection (DI) in a gasoline engine can deliver even higher fuel efficiencies that are on par with diesel engines. The VVA/DI engine can become strategically important to America and other countries dependent on a gasoline-based transportation economy. 
     Likewise, the use of a continuously variable valvetrain for the intake side of a gasoline engine coupled with an Early Intake Valve Closing (EIVC) load control strategy can significantly lower engine-out NOx emissions by lowering an engine&#39;s effective compression ratio at light and moderate engine loads. 
     What is needed in the art is a simplified, inexpensive, and reliable system for varying the lift, duration, and timing of engine valves which employs relatively few moving parts and affords a relatively small packaging envelope in the engine compartment of a vehicle. 
     It is a principal object of the present invention to reduce the complexity, cost of manufacture, and difficulty of manufacture of a system for varying the lift, duration, and timing of engine valves. 
     It is a further object of the present invention to reduce the packaging envelope required for such a system relative to prior art systems. 
     SUMMARY OF THE INVENTION 
     Briefly described, a HELP system in accordance with the present invention defines a mechanical VVA device for scheduling poppet combustion valve lift events on an internal combustion engine. Designed for ease of manufacture and reduced cost, the device varies valve lift, duration, and phasing in a dependent manner for one or more banks of engine valves. Using a single electrical rotary actuator per bank of valves to control the VVA device, the lift events can be varied for either or both the exhaust or intake valves, depending on how many such systems are employed. The valve actuation energy comes from a conventional engine camshaft that is driven by a belt or chain. The controlling actuator, which may be powered electrically, may receive its energy from the engine&#39;s alternator. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The present invention will now be described, by way of example, with reference to the accompanying drawings, in which: 
         FIG. 1  is an isometric view showing an exemplary HELP system, in accordance with the present invention, on a dual valve assembly at a representative camshaft single lobe position; 
         FIG. 2A  is a cross-sectional elevational view of the HELP system in a high engine load mode, showing the rocker roller on the base circle portion of the input camshaft; 
         FIG. 2B  is a cross-sectional elevational view of the HELP system in a high engine load mode, showing the rocker roller on the nose portion of the input camshaft; 
         FIG. 3A  is a cross-sectional elevational view of the HELP system in a low engine load mode, showing the rocker roller on the base circle portion of the input camshaft; 
         FIG. 3B  is a cross-sectional elevational view of the HELP system in a low engine load mode, showing the rocker roller on the nose portion of the input camshaft; 
         FIG. 4  is a family of representative cam timing, lift, and duration curves for a HELP system in accordance with the present invention; and 
         FIG. 5  is an exploded isometric drawing of a lifter adjuster assembly for use in the HELP system shown in  FIGS. 1-3 . 
     
    
    
     Corresponding reference characters indicate corresponding parts throughout the several views. The exemplification set out herein illustrates one preferred embodiment of the invention, in one form, and such exemplification is not to be construed as limiting the scope of the invention in any manner. 
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Full Engine Load Operation 
     In  FIGS. 1-3 , a HELP VVA system  100  in accordance with the present invention is shown at one of the inner cylinder locations of a naturally aspirated, inline 4-cylinder gasoline engine. The combustion chambers are typically of the pent roof variety, with four valves per cylinder (two intake, two exhaust). HELP system  100 , when applied to intake valves, manages an engine&#39;s intake gas exchange process with changes in the angular position of control shaft  1 . In  FIG. 2A  and  FIG. 2B , HELP system  100  is shown in a high engine load mode, and in  FIG. 3A  and  FIG. 3B , the HELP system is shown in a low engine load mode. In each of these pairs of figures, a view of the mechanism with the input roller  10  on the base circle portion  40  of cam lobe  6  appears to the left (FIGS.  2 A, 3 A), and a similar view with the input roller  10  on the nose portion  38  of cam lobe  6  (point of maximum lift) appears to the right (FIGS.  2 B, 3 B). Also, note that the front input cam bearing  3 , the front control shaft pivot housing  28 , lash spring  39 , and front output cam bearing  21 , as shown in  FIG. 1 , have been omitted for clarity in explaining the operation of HELP system  100 . 
     Referring to  FIG. 1 , axis of rotation  9  of control shaft  1  is centrically disposed in control shaft pivot housing  28  of internal combustion engine  62 , for rotation of control shaft  1 . Referring now to FIGS.  2 A, 2 B, control shaft  1  is eccentrically fixed to eccentric control shaft disc  14  such that disc  14  rotates eccentrically about axis of rotation  9  when control shaft  1  is rotated. High engine load events as shown in FIGS.  2 A, 2 B are produced by the mechanism whenever control shaft  1  is rotationally positioned such that the input rocker pivot center  4  (which is also the geometric center of disc  14 ) and output cam pivot center  5  are coincidental. At each engine cylinder is cam lobe  6  integral to nodular input camshaft  2 , centered axially between two engine valves  7 , 8  for a given cylinder. A rocker roller  10 , formed preferably of hardened steel, is free to rotate about a steel pin  11  staked in place within the input rocker clevis  12 . As input camshaft  2  rotates clockwise, the opening flank  25  of cam lobe  6  pushes rocker roller  10  upward conventionally, causing input rocker subassembly  13  to rotate in a clockwise direction about disc  14  and about input rocker pivot center  4 . As rocker subassembly  13  rotates, it turns about pivot center  4  of control shaft disc  14  via needle bearing  15 . 
     Integral to input rocker subassembly  13  is a foot  16 , formed preferably of cast steel, protruding from rocker bearing housing  17 . Foot  16  engages an output cam shoe  18  contained in a boss feature  19  at the center of output cam subassembly  20 , also formed preferably of cast steel. As input rocker subassembly  13  pivots clockwise about pivot center  4 , foot  16  forces output cam subassembly  20  to also rotate clockwise about the fixed output cam pivot center  5 . Rotation of output cam subassembly  20  is facilitated by front and rear output cam bearings  21 , 22  contained within output cam body  23 . The fixed output cam pivot center  5  is located by cast inner races  24 , 27  of the control shaft pivot housings  28  that are bolted to cylinder head  60  of engine  62 . In  FIG. 1 , two cast inner races disposed between inner races  24  and  27  are concealed by spring  39  and output cam subassembly  20 . 
     Note that in the high engine load mode of control shaft  1 , the input rocker and output cam subassemblies  13 , 20  move in unison on identical centers, that is to say, there is no relative motion between input rocker foot  16  and output cam shoe  18 . Hence, in  FIG. 2B , input cam lobe  6 , input roller  10 , input rocker subassembly  13 , and output cam subassembly  20  act as a simple four-bar linkage, with respect to the virtual link that exists between fixed input camshaft pivot center  29  and fixed output cam pivot center  5 . 
     Clockwise rotation of output cam subassembly  20  advances output cam profiles  30 , 31 , ground into the output cam body  23 , to where the radius of the output cam increases beyond that of the base circle portion  32  of the cam profile. The further that output cam subassembly  20  is rotated about the control shaft pivot housing inner races (only races  24  and  27  are visible), the greater the lift imparted through the finger follower rollers  33 . The right end of each finger follower  34  pivots about the ball-shaped tip  35  of a conventional hydraulic valve lash adjuster  36  mounted in cylinder head  60 . Pushing down on the centrally-located finger follower rollers  33  transmits lift to engine valves  7 , 8  via pallet  37  at the left ends of the finger followers  34 . 
     Although the preferred embodiment described herein is depicted with low friction roller finger followers, using roller as a contact feature between output cam assembly  20  and finger follower  34 , HELP system  100  is not limited to this type of a standard valvetrain. Another embodiment within the scope of the present invention may have crowned bucket type tappets (not shown here) with slightly different-shaped output cam profiles (not shown) as are known in the art. 
     When eccentric control shaft disc  14  is in the high engine load mode, as shown in  FIGS. 2A and 2B , maximum lift is imparted to engine valves  7 , 8  whenever rocker roller  10  reaches the nose portion  38  of cam lobe  6 . At this point, the input rocker and output cam subassemblies  13 , 20  cease to move in the clockwise direction. As input cam lobe  6  rotates further in the clockwise direction, nose portion  38  of lobe  6  slips past rocker roller  10 , and lash spring  39  forces the output cam and input rocker subassemblies  20 , 13  to rotate counter-clockwise. This counter-clockwise rotation, in turn, reduces lift produced between the output cam profiles  30 , 31  and the finger follower rollers  33 . Eventually, as camshaft  2  continues to rotate clockwise, rocker roller  10  reaches the base circle portion  40  of cam lobe  6  where lift remains at zero until the next engine event occurs for that cylinder. 
     The motion just described produces a peak lift profile similar to peak lift profile  41  shown in  FIG. 4 , to maximize gas flow to (intake valve) or from (exhaust valve) of engine  62 . 
     Low Engine Load Operation 
     Referring now to FIGS.  3 A, 3 B, control shaft  1  and the changing force load of the input rocker subassembly  13  are supported by bearings  42  ( FIG. 1 ). An electromechanical actuator (not shown) operationally connected to control shaft  1  can change the angular position of control shaft  1  and eccentric control shaft disc  14  about the center of bearings  42  to vary engine load. Bearings  42  preferably are needle-type bearings, although dry-type sleeve bearings (not shown) may be used in some applications. 
     Referring to  FIG. 2A  through  FIG. 3B , when control shaft  1  is rotated significantly clockwise relative to its high engine load mode position, HELP system  100  produces lower lift events (see region  43  in  FIG. 4 ) with reduced duration, corresponding to lower engine loads. When this happens (FIGS.  3 A, 3 B), input rocker pivot center  4  of control shaft disc  14  moves inward toward camshaft  2 , away from the fixed pivot center  5  location of output cam subassembly  20 . Thus, when input cam lobe  6  induces angular motion to the rocker subassembly  13 , relative rolling and sliding motion results between input cam foot  16  and output cam shoe  18 , since a second four-bar linkage is now created between the virtual ground link of input rocker pivot center  4  and output cam pivot center  5 , the virtual link between input rocker pivot center  4  and the rocker foot is curvature center  44 , the virtual link between foot curvature center  44  and the shoe curvature center  45 , and the virtual link between shoe curvature center  45  and output cam pivot center  5 . 
     Likewise, when control shaft assembly  1  is in the lowest engine load mode (FIGS.  3 A, 3 B), finger follower rollers  33  spend most of their path on the base circle portion  32  of output cam profiles  30 , 31 , just barely reaching the opening ramp  46  of the output cam profile, whenever the input rocker roller  10  is aligned with the nose portion  38  of cam lobe  6 . Thus, HELP system  100  can generate a short and shallow lift event (curve  47  in  FIG. 4 ), suitable for the lightest of all engine loads. 
     It will be observed that displacement of the control shaft position from that shown in FIGS.  2 A, 2 B to that shown in FIGS.  3 A, 3 B serves a) to advance the position of input roller  10  on cam lobe  6 , thereby advancing the start of valve opening, and b) to advance the contact point of profile  30  or  31  with finger roller  33 , thereby reducing the potential valve lift. Thus, varying the angular position of control shaft  1  between the high engine load position illustrated in FIGS.  2 A, 2 B and the low engine load position just described for FIGS.  3 A, 3 B produces the entire lift curve family depicted in  FIG. 4 . 
     Special Features 
     One novel feature of HELP system  100  is elimination of relative motion between input rocker foot  16  and output cam shoe  18  when the control shaft is in the high engine load lift mode. Since both rocker pivot center  4  and output cam pivot center  5  are coincidental at high load lift mode, there is also no relative angular motion between the foot  16  of input rocker subassembly  13  and shoe  18  of output cam assembly  20 , respectively). This makes the high load lift cam profile design fairly straight forward from a kinematic standpoint. 
     A second important improvement is the convenient lift adjuster as shown in  FIG. 5 . In a prior art EIVC throttleless-load control scheme, small lift variations from one cylinder to the next can substantially affect engine stability at light loads such as idle. For instance, at a light load, a global lift command of 2 mm might be issued by the engine controller. If a prior art VVA system delivers the correct 2 mm valve lift profiles to three of the cylinders, but a 0.5 mm lift error occurs at the fourth cylinder, a significant gas flow error will result. However, the same 0.5 mm lift error at a full load lift of 10 mm in a conventionally-throttled system would have little noticeable effect at the same idle load. 
     The lift adjuster shown in  FIG. 5  includes lift adjuster mechanism  104 , in an exploded view. The previously described output cam shoe  18  includes a threaded stud  48 , welded to a hardened, wear resistant pad  57 , ground to a precise curvature. Right hand external threads (not shown) on stud  48  are provided at X threads per inch. Correspondingly, the inner diameter of adjusting screw  50  also contains internal threads  49  of the same pitch. The outer diameter of adjusting screw  50 , however, has external right hand threads  51  with a pitch of preferably about X+1 threads per inch. Likewise, bore  52  in boss feature  19  of the output cam body  23  contains right handed, internal threads  53  that have the X+1 pitch, and jam nut  54  contains internal right handed threads of X pitch. 
     During the final assembly of the output cam, the diameter of stud  48  of shoe  18  is loosely inserted through one or more Belleville washers  55  into larger bore  52  in boss feature  19  of output cam body  23 . Next, the internal  49  and external  51  threads of adjusting screw  50  are simultaneously engaged over threaded stud  48  of shoe  18  and into internal threads  53  of boss feature  19 , respectively. Since the two pitches vary by only one thread per inch, several turns of adjuster screw  50  are required to seat output cam shoe  18  fully against Belleville washers  55  into output cam body  23 . Flats  56  machined into output cam body  23  mate with flat sides  58  of pad  57  of output cam shoe  18  to keep it from turning as Belleville washers  55  are loaded. Lastly, jam nut  54  is screwed on over threaded stud  48  of output cam shoe  18  and tightened against adjusting screw  50  to lock the adjuster against engine vibrations. 
     After input camshaft  2  and cam bearings  3  are installed on cylinder head  60  that has been bolted to engine  62 , the completed HELP system  100  is lowered onto the head and bolted down with fasteners (not shown) through bores  59  in control shaft pivot housings  28 . 
     A cylinder head  60  equipped with HELP system  100  provides an engine manufacturer with several options to balance the cylinder-to-cylinder gas flow. The HELP system lift adjustment provision provides a unique flexibility to choose the best method. Gas flow can be adjusted either on an individual cylinder head in a flow chamber environment, or on a completed running engine. 
     Assembly line calibration typically occurs at an automated test stand, with either a precision air flow rate meter for calibrating individual completed cylinder heads, or with a bench type combustion gas analyzer for calibrating fully assembled engines. For balancing individual cylinder heads, lift can be adjusted either statically to match a desired steady-state, steady-flow-rate target with the camshaft fixed, or dynamically with the camshaft spinning, by measuring the time-averaged flow rate for each cylinder. However, HELP system  100  can also be adjusted dynamically in a repair garage with a running engine, using cylinder-to-cylinder exhaust gas analysis techniques with a portable fuel-to-air ratio analyzer. 
     Referring again to  FIG. 1 , to make a lift adjustment, jam nut  54  is loosened with a first wrench while holding adjusting screw  50  in position with a second wrench. Once jam nut  54  is loosened 2-3 revolutions, the second wrench can be used to adjust the lift simultaneously at engine valves  7 , 8 . 
     Relative to the contact face of foot  16 , adjusting screw  50  is rotated counter-clockwise to increase lift. This causes adjusting screw  50  to pull away from boss feature  19  of output cam subassembly  20  and input camshaft  2 . Since output cam shoe  18  is constrained from rotating by the flats  56  machined into the output cam body, it is also pushed away from adjusting screw  50 . But the difference in thread pitches causes adjusting screw  50  to pull away from output cam subassembly  20  more slowly than output cam shoe  18  is pushed away from adjuster screw  50 , ultimately causing output cam shoe  18  to be pushed farther away from output cam subassembly  20 . If the engine is not running and the rotary position of control shaft  1  is fixed, the resultant motion of the output cam shoe  18  with respect to output cam subassembly  20  causes output cam subassembly  20  to rotate clockwise relative to the input rocker subassembly  13 . This in turn increases lift at valves  7 , 8 . 
     However aggressive the cam profiles  30 , 31  are, a careful selection of the threaded pitch “X” in the adjuster parts can yield as little as a 100 micron lift change at valves  7 , 8  for every revolution of adjuster screw  50  for each of the individual cylinders of engine  62 . This is an ideal flow adjustment resolution for balancing the gas flow across all the cylinders. 
     While the invention has been described by reference to various specific embodiments, it should be understood that numerous changes may be made within the spirit and scope of the inventive concepts described. Accordingly, it is intended that the invention not be limited to the described embodiments, but will have full scope defined by the language of the following claims.