Adjusting valve timing to deactivate engine cylinders for variable displacement operation

An internal combustion engine having two cylinder banks and adjustable camshaft timing is disclosed in which the camshafts in one cylinder bank are adjusted so that there is no net flow from the cylinders to effectively disable the cylinder bank. In particular, exhaust valve timing is advanced so that the maximum valve lift occurs approximately at bottom center between expansion and exhaust strokes and intake valve timing is advanced so that maximum valve lift occurs approximately at bottom center between intake and compression strokes. Also disclosed is an engine in which an intake and an exhaust camshaft on a single bank are coaxial with valve timings adjusted by rotating the inner of the two camshafts with respect to the outer of the two camshafts.

BACKGROUND

1. Technical Field

The present disclosure relates to using a variable camshaft timing device to deactivate a portion of engine cylinders.

2. Background Art

Variable camshaft timing (VCT) devices are used to change the phase relationship between the crankshaft and the camshaft lobes that control actuation of the intake and exhaust poppet valves. By controlling the phase relationship, the fuel economy, performance, and emissions can be improved by providing less valve overlap at low engine speed operation to improve combustion stability, and more valve overlap at higher engine speed to improve engine power and performance.

It is known that by deactivating a portion of engine cylinders when the engine is operating at light load, that overall fuel economy can be improved. Typically, cylinders are deactivated by deactivating intake and exhaust poppet valves. In U.S. Pat. No. 6,237,559 B1, commonly assigned to the assignee of the present application, cylinder deactivation is provided by disabling exhaust poppet valves. The intake valves are not disabled, but are opened and closed generally symmetrically about a top center or bottom center position of the piston. This results in a cost savings because valve deactivators are provided for exhaust valves and not intake valves. However, such a system does not eliminate valve deactivators, which are known to be problematic. For example, there are typically issues associated with latching the valve deactivators and difficulties in ensuring that they are latched within one engine cycle, particularly at high engine speeds.

In U.S. Pat. No. 5,642,703 commonly assigned to the assignee of the present application, cylinder deactivation is provided by adjusting a single camshaft which actuates both the intake and the exhaust valves, or alternatively, using dual camshafts. However, this reference teaches that both camshafts may be linked together with one phase and that adjusting the timing of the valve lift events has no effect on the relative timing between the exhaust valve lift event and intake valve lift event, i.e., the timing between exhaust and intake valve lift events remains constant regardless of phase shifting. Because the intakes and exhausts are not independently adjusted, one embodiment shows both valve events retarded such that the exhaust valve event is roughly centered about top center (TC) of piston movement between the exhaust and intake strokes and the intake valve event is roughly centered about bottom center (BC) between the intake and compression strokes. The valve events are optimized for normal operation, not for deactivated operation. Thus, it is not possible to adjust both intake and exhaust valve events such that they are symmetrical about a rotational position of the crankshaft when the direction of the piston changes, i.e., TC or BC. Also, valve overlap, which lasts about 25 degrees of crank rotation is not adjusted in such a system because the intake and exhaust valve events move in lock step. It is typical for the intake system to operate at a vacuum and the exhaust system to be at a pressure slightly higher than atmospheric. Thus, during the valve overlap period while the valves are in the deactivated condition, exhaust gases flow into the intake and are then inducted by the activated cylinders. While there may be operating conditions where such flow from exhaust to intake is advantageous, it is desirable to control such flow. For example, when operating at low torque conditions, there is a considerable intake manifold vacuum to drive flow from the exhaust to the intake. However, this is an operating condition where no additional exhaust gases can be accommodated without impairing combustion stability. At higher torque conditions, exhaust gases can be accommodated in the combustion charge, but there is little manifold vacuum to drive the flow from exhaust to intake. While the strategy for valve deactivation disclosed in U.S. Pat. No. 5,642,703 is suitable for many applications, various advantages may be provided by appropriate control of valve overlap that cannot be achieved with such a strategy.

SUMMARY

According to an embodiment of the present disclosure, a portion of cylinders are deactivated by opening and closing exhaust valves generally symmetrically about a bottom center position of the piston and opening and closing intake valves generally symmetrically about a bottom center position of the piston with the intake and exhaust camshafts adjusted by a mechanical adjustment device to provide substantially no valve overlap. By opening and closing the intake or exhaust valves in such a manner, the gases inducted into engine cylinders while the piston is moving toward bottom center are exhausted from engine cylinders while the piston is moving upward from bottom center. By causing the valve event to be generally symmetrical about bottom center, the quantity of gases entering and exiting the cylinder is substantially equal. Thus, there is substantially no net flow of gases.

Valve deactivation according to the present disclosure presents an advantage by using exhaust cam phasing to effectively disable the cylinder as opposed to the solution shown in U.S. Pat. No. 6,237,599 in which exhaust valve deactivators are employed. As typical of hydraulic variable cam timing devices, the VCT device in '599 has insufficient range of authority to adjust exhaust valve timing to provide substantially no valve overlap and relies on valve deactivators on the exhaust valves. The valve deactivation strategy of the present disclosure also presents an advantage over strategies such as shown in U.S. Pat. No. 5,642,703 because adjusting the exhaust and intake valve events independently allows operation with substantially no valve overlap. Thus, unintended exhaust flow into the intake is substantially eliminated.

Use of a mechanical cam phasing device, such as disclosed herein, provides a greater range of authority, thereby facilitating adjusting the exhaust camshaft to a position in which there is substantially no net flow of gases through the exhaust valves. Such camshaft phasing device acts upon each camshaft as a whole. Thus, in an engine having two cylinder banks, one bank of cylinders may be operated with conventional valve timings, whereas, the valve timings of the other bank of cylinders are adjusted to deactivate the cylinders.

In deactivator systems in which valve deactivators are provided to cause the valves to stay closed, the gases trapped in the cylinder are compressed and expanded during each engine revolution. After several revolutions, the average pressure in the cylinder is approximately atmospheric with pressures exceeding atmospheric when the piston is near top center and pressures below atmospheric when the piston is near bottom center. During periods of vacuum in the cylinder, oil from the cylinder walls is drawn past the piston ring(s) and into the combustion chamber. When such cylinders are reactivated, the excess oil is burned, which is generally undesirable. Thus, it is desirable to avoid such a vacuum existing in the cylinder. According to an aspect of the present disclosure, by having an exhaust valve or an intake valve open during the period when the piston is near bottom center, vacuum that would otherwise develop in the cylinder cannot develop. Instead, according to an embodiment of the present disclosure, the pressure in the cylinder is substantially the same as that in the exhaust manifold when the exhaust valve opens and substantially the same as the pressure in the intake manifold when the intake valve opens thereby relieving any vacuum that would otherwise build up due to piston movement. Because substantially no vacuum develops in the cylinders, oil pumping past piston rings is mitigated according to an embodiment of the present disclosure.

Furthermore, most valve deactivation systems deactivate (or activate) valves in an abrupt manner causing one-half of engine cylinders to turn off (or turn on) abruptly. It is difficult to manage engine torque to obtain a smooth transition under such conditions. Yet another advantage of an embodiment of the present disclosure is that the valve adjustments to the symmetric valve timing can be accomplished over a number of engine cycles to facilitate a smooth transition in torque.

DETAILED DESCRIPTION

As those of ordinary skill in the art will understand, various features of the embodiments illustrated and described with reference to any one of the Figures may be combined with features illustrated in one or more other Figures to produce embodiments that are not explicitly illustrated or described. The combinations of features illustrated provide representative embodiments for typical applications. However, various combinations and modifications of the features consistent with the teachings of the present disclosure may be desired for particular applications or implementations. The representative embodiments used in the illustrations relate generally to a multi-cylinder, internal combustion engine having a non-hydraulic variable cam timing device to vary the angular relationship between the camshaft and crankshaft and/or between sets of camshaft lobes to provide variable displacement operation by selectively deactivating cylinders by adjusting valve timing. However, those of ordinary skill in the art may recognize similar applications or implementations with other engine/vehicle technologies.

FIGS. 1-3illustrate representative arrangements for controlling the rotational phase relationship between two or more rotating shafts of an internal combustion engine using a device or devices according to the present disclosure. Those of ordinary skill in the art will recognize that although the representative examples ofFIGS. 1-3illustrate a device according to the present disclosure used to change the relationship between two camshafts and/or between a camshaft and a crankshaft to selectively deactivate one or more cylinders or a bank of cylinders, other suitable applications for such a device exist. Likewise, a mechanical phaser according to the present disclosure is not limited to the illustrated arrangements and one or more such devices may be used depending upon the particular application and implementation. In the representative arrangements illustrated, the function or role of any particular camshaft may vary by application and implementation to operate intake valves, exhaust valves, or both.

In the arrangement represented by the top view ofFIG. 1Aand the front view ofFIG. 1B, an engine having a single cam axis for each cylinder head includes a coaxial camshaft having an inner exhaust camshaft22for operating exhaust valves and an outer intake camshaft24for operating intake valves, or vice versa. As such, an in-line or I-type cylinder configuration would include a single crankshaft20and coaxial camshaft22,24whereas a V-type configuration would include two cylinder heads each having a coaxial camshaft22,24connected to a single crankshaft20. In either arrangement, outer camshaft24may be driven by a drive sprocket26connected by a chain or belt28to crankshaft20. Alternatively, outer camshaft24may be driven by a hydraulic phaser (not shown) to selectively change the phase relationship between crankshaft20and outer camshaft24, although use of a hydraulic phaser in combination with a mechanical phaser according to the present disclosure may not achieve all of the advantages of exclusive use of mechanical phasers, such as cold temperature performance, lower oil pressure operation, range of control to provide substantially no valve overlap, etc. A mechanical phaser30, as described in greater detail herein, drives inner camshaft22using first and second oppositely biased worm assemblies to selectively change the phase relationship between outer camshaft24and inner camshaft22.

The representative arrangement illustrated inFIG. 2represents a dual overhead camshaft arrangement with two cam axes per cylinder head. A first cam axis is associated with first inner camshaft22′ and first outer camshaft24′ that both operate intake valves with at least two intake valves per cylinder. A second cam axis is associated with second inner camshaft22″ and second outer camshaft24″ that both operate exhaust valves with at least two exhaust valves per cylinder. Outer camshafts24′,24″ may be driven directly from chain28′ or indirectly through conventional hydraulic phasers (not shown) positioned at the front of each coaxial camshaft. Mechanical phasers30′,30″ control timing of the valves operated by cam lobes on corresponding inner camshafts22′,22″. Appropriate staggered or offset control of the valve opening timing of the two valves provides a longer overall duration of the opening event.

A mechanical phaser according to the present disclosure has no inherent limits to its range of control as compared to hydraulic phasers that are typically limited to around thirty degrees of total motion at the camshaft (or sixty degrees as measured by crankshaft rotation). As described in greater detail herein, this range of control facilitates generally symmetric valve timing relative to cylinder bottom dead center to provide substantially no overlap during cylinder deactivation for variable displacement engine applications. Hydraulic phaser arrangements may be used to implement dual dependent variable cam timing with a first drive chain coupled to the crankshaft to drive one of the camshafts through a first phaser and a second drive chain running between the dual camshafts to drive the other camshaft through a second phaser. This provides the second camshaft a greater total range of adjustment relative to the crankshaft than what could be achieved with a dual independent configuration using conventional hydraulic phasers. The mechanical phasers of the present disclosure may be used to implement dual independent variable cam timing with the phaser of each camshaft receiving its input from the drive chain coupled to the crankshaft. This mechanical dual independent configuration has a larger range of authority over the second camshaft timing than does the hydraulic dual dependent arrangement, but does not force one of the phasers to carry the loads of both camshafts.

In the representative arrangement ofFIG. 3, crankshaft20″ drives sprocket32, which is coupled to a conventional one-piece camshaft40via mechanical phaser30′″. As best illustrated in the cross-section ofFIG. 14, drive sprocket32is located near the front bearing and the radial chain load from chain28″ is supported by camshaft40and associated bearings that carry the cantilevered load out past the front bearing. This arrangement uses mechanical phaser30′″ to selectively change the phase relationship between rotating crankshaft20″ and camshaft40.

FIGS. 4-6illustrate a representative coaxial camshaft for use with a device for changing phase relationship between first and second rotating shafts according to the present disclosure. In this arrangement, outer camshaft70carries adjustable cam lobe76(best shown inFIG. 6) and fixed cam lobe78(best shown inFIG. 5) on one diameter to ensure concentricity and carries associated radial loads of valve actuation. Inner camshaft72carries minimal or no bending load from the valve opening forces, but provides the torque needed to drive adjustable cam lobes76. Inner camshaft72is supported on two bushings (not shown) at either end. Fixed cam lobes78are rigidly connected to outer camshaft70with set screw80, or camshaft70may be swaged into the fixed cam lobes or attached by other means. Adjustable cam lobes76are free to rotate on outer camshaft70via a bushing74within their inner diameter. Adjustable cam lobes76are driven by inner camshaft72via dowel pin82that connects through slots in outer shaft70. During operation, a variable cam timing device as illustrated and described herein may be used to change the phase relationship between outer camshaft70and inner camshaft72to effect symmetric valve timing with substantially no valve overlap for cylinder deactivation during variable displacement operation.

FIGS. 7-12illustrate one embodiment of a device for selectively varying a relative angular position between first and second rotating shafts of an internal combustion engine according to the present disclosure. Device100includes a first worm assembly102coupled to first rotating shaft110via worm carrier160and second rotating shaft112via gear132, and torsionally preloaded to provide a first torque tending to advance the relative angular position of the second shaft relative to the first shaft, and a second worm assembly104coupled to the first and second rotating shafts110,112and torsionally preloaded opposite to the first assembly102to provide a second torque simultaneously tending to retard the relative angular position of the second shaft112relative to the first shaft110such that backlash and torsional free play between the two shafts is removed and the relative angular position of the first and second shafts remains substantially constant during rotation of the shafts whenever their angular phase relationship is not being adjusted. At least one actuator acts on rear side gear116and front side gear118and selectively applies an actuating torque to respective worm pinions120,122of the first and second worm assemblies102,104to rotate the second shaft112relative to the first shaft110and change the relative angular position between the first and second shafts110,112.

In one embodiment, device100is fitted to the rear of a coaxial camshaft having an outer camshaft110and inner camshaft112. A front side plate130is electron beam welded onto the rear end of outer camshaft110. A worm gear132is welded onto a tube134disposed within outer camshaft110with a running fit, and is pressed tightly onto a reduced diameter section136toward the rear end of inner camshaft112. A screw140engages corresponding threads (not shown) within inner camshaft112and clamps washer142and tube134against the shoulder formed by reduced diameter portion136of inner camshaft112. Additional torque carrying capacity may be provided by a splined interface or by a pin (not shown) passing through tube134and inner camshaft112. Depending on the particular application and implementation, a common pin could be used to attach a cam lobe76and tube134to inner camshaft112.

As best illustrated in the cross-sections ofFIGS. 8-9, a worm carrier assembly150contains first and second worm assemblies102,104. Carrier assembly150includes carrier160that is assembled around worm gear132which is attached to inner camshaft112through tube134. Four flat head screws or rivets162,164,166,168hold the two halves of carrier160together to contain the worm assemblies102,104. Four additional flat head screws or rivets182,184,186,188pass through rear side plate190, through carrier160, and into front side plate130, attaching the carrier assembly150to outer camshaft110via the electron beam weld previously described. The actuator, which in this embodiment includes front side gear118and rear side gear116, includes teeth in meshing engagement with worm pinions120,122and are captured between carrier160and front side plate130and rear side plate190, respectively, in a manner that constrains fore/aft, and radial motion of side gears116,118, but allows each side gear116,118to rotate about the axis of camshafts110,112, while transferring torque to/from worm pinions120,122. The meshing engagement between side gears116,118and worm pinions120,122is best illustrated in the cross-section ofFIG. 10.

In the embodiment shown inFIGS. 7-12, the axes of worm pinions120,122and the common axis of side gears116,118do not intersect with one another. As such, worm pinion gears120,122and the mating side gears116,118are constructed with a hypoid design, similar to the gears used in the rear axle of many rear wheel drive vehicles with front mounted engines. Worm pinions120,122have a running fit to respective worm shafts206,210and transmit torque to a quill clamp220that has a blade224engaged into a slot226on the end of pinion gear120as best illustrated in the cross-sections ofFIGS. 10-12. This quill clamp220, in turn, transmits torque to a torsion element implemented by a quill230in this embodiment. Quill230passes through the length of worm shaft206, and the torque is received by another quill clamp220′ at the opposite end of worm shaft206. Each quill clamp is attached to quill230by a pair of set screws236. A torsional preload can be imparted to quill230by twisting and holding the ends of the quill that extend beyond clamps220and220′ while the set screws are being tightened. This second quill clamp220′ also has a blade224′ that transmits the torque to drive lugs240′,242′ on the corresponding end of worm shaft206. When quill230reaches a certain level of torsional load, elastic deformation of quill230allows quill clamp220to rotate relative to worm shaft206so that blade224contacts drive lugs240,242of worm shaft206at the pinion end and any increased level of torque is transmitted from pinion gear120to worm shaft206, at that contact, without further deformation of quill230. A worm260is secured to worm shaft206so that rotation of worm shaft206produces rotation of worm260, and consequent rotation of worm gear132such that inner camshaft112rotates relative to worm carrier assembly150and outer camshaft110.

In operation, an actuator under control of an engine control module to deactivate an associated group of cylinders applies a frictional dragging force on front side gear118causing it to rotate backward relative to outer camshaft110and worm carrier assembly150. The relative motion causes worm pinions120,122to rotate respective worms260,260′ in a clockwise direction as viewed from the worm pinion gear end of each worm assembly. Because worms260,260′ both have their teeth oriented in right hand helices, clockwise rotation of worms260,260′ will cause inner camshaft112to advance or move in the direction of rotation indicated by arrows270relative to outer camshaft110. Similarly, when an actuator under control of the engine control module applies a frictional dragging force to rear side gear116causing rear side gear116to rotate backward relative to outer camshaft110and worm carrier assembly150, the relative motion causes counterclockwise rotation of worms260,260′ to retard inner camshaft112relative to outer camshaft110.

During operation of the engine, as the outer camshaft110and inner camshaft112rotate to actuate valves coupled by cam followers and corresponding cam lobes, the action of the cam lobes on the followers produces an oscillating torsional load between outer camshaft110and inner camshaft112. As such, it is desirable to reduce or eliminate backlash within the device to prevent noise and wear and to maintain a desired phase relationship between the rotating shafts when no actuating torque is applied to rear side gear116or front side gear118. According to one aspect of the present disclosure, this is accomplished by having opposite hand torsional preload on the two worm assemblies102,104. When the variable cam timing device is rotating at a steady-state without advancing or retarding, worm assembly102is torsionally preloaded or biased in a direction attempting to advance inner camshaft112, while worm assembly104has an opposite preload or bias that attempts to retard inner camshaft112. Because worms260,260′ have a relatively fine tooth pitch, a torque applied to the worms can produce rotation of worm gear132. However, the converse does not produce movement, i.e. because of the friction locking between worm gear132and worms260,260′, a torque applied to worm gear132will not produce rotation of worms260,260′, as long as the worms have any amount of resisting torque. The torsional preloads between worm pinion gears120,122and the associated worm shafts206,210are transmitted through the corresponding quills230,230′, respectively. In the steady state condition, worm assembly102has a clockwise preload on its quill230, and worm assembly104has a counterclockwise preload on its quill230′.

FIG. 13illustrates an alternative embodiment of a carrier assembly150′ having intermediate gear assemblies300,302, with one end in meshing engagement with a corresponding worm pinion gear120′,122′, respectively, and an opposite end in meshing engagement with the actuating gear(s) implemented by a front side gear and rear side gear as described with respect to the embodiment illustrated inFIGS. 7-12. Intermediate gear assemblies300,302are positioned with axes intersecting the axes of corresponding pinion gears120′ and122′, and also the common axis of the front and rear side actuating gears such that bevel gears may be used rather than hypoid gears.

FIG. 14is a cross-sectional view of a mechanical phaser30′″ used with a conventional camshaft40such as illustrated in the diagram ofFIG. 3. Camshaft40includes cam lobes318to operate corresponding intake/exhaust valves and phaser30″ may be constructed with no mechanical limits on its range of phase adjustment relative to camshaft drive sprocket32. Camshaft drive sprocket32may be attached in any suitable manner to rear side plate190′, or may be integrally formed therewith. Outer stub shaft134′ is secured to rear side plate190′ and rotates with sprocket32, which is located near front bearing324. The radial chain load applied through sprocket32is supported by camshaft40, which is in turn supported by front bearing324as well as by various other bearings, such as second camshaft bearing326to carry the cantilevered load of the mechanical phaser and sprocket assembly out past front bearing324. During operation, a first axial actuator320selectively applies a frictional dragging torque to rear side gear116′ to retard the rotation of camshaft40relative to sprocket32and crankshaft20″ (FIG. 3) as previously described with reference to the embodiment ofFIGS. 7-13. Similarly, to advance rotation of camshaft40relative to sprocket32and crankshaft20″ (FIG. 3), a second axial actuator322selectively applies a frictional dragging torque to front side gear118′. First and second actuators or brakes320,322may be mounted to the engine front cover or other stationary component depending on the particular application and implementation.

FIG. 15is a cross-sectional view of another arrangement for selectively changing the phase relationship between two or more rotating shafts of an internal combustion engine according to the present disclosure. In the arrangement ofFIG. 15, a first (front) device30′″ controls the phase relationship of inner coaxial camshaft112′ relative to drive sprocket32while a second (rear) device100′ controls the phase relationship of outer coaxial camshaft110′ relative to inner coaxial camshaft112′. Similar to the arrangement illustrated inFIG. 14, the arrangement ofFIG. 15eliminates the use of any hydraulic phaser, which may facilitate use of a smaller oil pump and lower operating oil pressure, resulting in increased fuel efficiency.

Referring now primarily toFIGS. 16-17, charts illustrating operation of a device for changing phase relationship between two rotating shafts of an internal combustion engine are shown. The charts ofFIGS. 16-17illustrate how an actuating torque provided by dragging forces applied on the front or rear side gears116,118is translated into torque on the torsionally preloaded worm assemblies102,104to rotate worm gear132. The horizontal axes illustrate a representative actuating or frictional drag torque applied to the side gears to advance or retard the relative rotation of the rotating shafts. The vertical axes illustrate representative values for the resulting torque on the worm assemblies102,104tending to advance or retard inner camshaft112relative to outer camshaft110. At all values of drag torque applied to a side gear, the sum of the torques resulting in the two worm assemblies102,104is equal to the torque applied to the side gear times the effective gear ratio between the side gear and the worm pinions. The amount of torque carried by one worm assembly as compared to the other worm assembly, however, changes because of differences in torsional preload torque, as well as changes in the elasticity within each worm assembly between the worm pinion and its worm.

In the embodiment illustrated in the chart ofFIG. 16, the quill assemblies are preloaded so that the torsion element, implemented in the representative embodiment by quills230,230′, will not be subjected to a reversal of torque direction. In contrast, for the embodiment illustrated in the chart ofFIG. 17, the quill assemblies have a torsion element with a lower preload that subjects the torsion element to a reversal of torque direction.

In the chart ofFIG. 16, line400represents the relationship between torque applied to worm assembly102, and line402represents the relationship between torque applied to worm assembly104as a function of frictional drag torque applied to front side gear118or rear side gear116. As illustrated in the charts ofFIGS. 16-17, moving toward the right hand side of the chart represents an increasing frictional drag or actuating force on the front side gear118, while moving toward the left represents an increasing frictional drag on the rear side gear116. Moving upward represents the resulting torque on the worm assemblies attempting to advance inner camshaft112, and moving downward represents torque that tries to retard inner camshaft112. When neither side gear116,118has a frictional drag or actuating force, the torque applied to the two worms260,260′ through the preloaded quills230,230′, respectively, balance each other. Worm assembly102is biased toward advancing worm gear132, while worm assembly104is biased toward retarding worm gear132. This neutral position is indicated at406.

During actuation, when a frictional drag of up to 0.5 Newton-meters (N-m) is applied to front side gear118as indicated in the region between406and408, the front side gear118and both worm pinion gears120,122rotate relative to carrier160. The quill230of worm assembly102sees an increasing clockwise torque, and quill230′ of worm assembly104sees a decreasing counterclockwise torque. Because both worm assemblies use quills of identical elasticity, the rate of torque increase in worm assembly102is equivalent to the rate of negative torque decrease in worm assembly104. At the drag torque level indicated at408, blade element224of the quill clamp220rotates to the point of contacting drive lugs240,242of the worm shaft206.

Once contact has been established between blade224and lugs240,242of worm assembly104, represented at408, the torsional stiffness of the worm shaft206prevails over the elasticity of the quills. Additional dragging torque at the front side gear, represented by the transition from408to410, is resisted by reduction of negative torque through the now stiff worm assembly104. Rotation of the worm pinions is minimal and consequently the additional torque imposed upon worm assembly102is minimal.

As the frictional drag on the front side gear increases from410to412, worm assembly104transitions from negative torque to positive torque, and the backlash in the gear meshes allows some rotation of the front side gear that does not impose any torque load through worm assembly104. This small rotation of the front side gear, however, does cause an additional load to be imposed upon the quill230of worm assembly102.

Up until this point, the dragging torque on the front side gear has not caused any rotation of the worms260,260′, nor any change of the angular orientation between the outer camshaft110and the inner camshaft112. However, any increase of dragging torque on the front side gear beyond412will impose an advancing torque on worm gear132from both of the worm assemblies102,104. With a substantial increase of dragging torque on the front side gear, represented onFIG. 16by414, worm assembly104, driving through the stiff worm shaft206, will generate substantial torque on its worm260′ while all of the torque on worm assembly102is passed through its elastic quill, and thus is limited to a lower maximum value. In effect, the stiffness of the worm shaft of worm assembly104allows generation of the “muscle” to advance the inner camshaft112, while the elasticity of the quill230provides sufficient torque to keep its worm260out of the way.

During normal activation of the phaser, it is usually desired to have a rapid adjustment of the phase relationship between the outer and inner camshafts,110and112respectively. The dragging force applied to the front or rear side gears130,190would be much larger than the range of loads shown onFIG. 16. The maximum loads may be determined by the durability strength of the phaser assembly100and would be controlled by limiting the frictional drag torques applied to the front and rear side gears116,118.

When both worm assemblies102,104see a clockwise torque as generally indicated in area414ofFIG. 16, an advancing torque is applied to worm gear132. In some applications of phaser100, the dynamic torque of the camshaft associated with operation of the intake and/or exhaust valves may be much higher than the torque applied to worm gear132by worm assemblies102,104, and inner camshaft112may be able to advance only during the portion of the dynamic torque load that is already trying to advance inner camshaft112. At other moments, when the dynamic load of the cam action is trying to retard inner camshaft112, worm assemblies102,104may be unable to advance worm gear132, but would have enough of a mechanical advantage or friction locking to hold it from retarding.

The left hand side of the chart inFIG. 16represents an actuating force or frictional drag on rear side gear116and operates in a similar fashion as described above with the function/operation of worm assemblies102,104reversed to retard inner camshaft112.

The chart ofFIG. 17illustrates operation of an embodiment having a smaller torsional preload of quills230,230′ than the embodiment illustrated inFIG. 16. Line500represents worm assembly torque as a function of actuating force for worm assembly102, while line502represents worm assembly torque as a function of actuating force for worm assembly104. With a smaller torsional preload, the torsion elements implemented by quills230,230′ are subjected to a reversal in torque direction before the corresponding quill clamp blades224contact the worm shaft drive lugs240,242at the pinion end of the worm shafts. Depending upon the particular application and implementation, the configuration illustrated in the chart ofFIG. 17may allow the device to be more responsive and provide faster advance and retard speeds than the embodiment illustrated in the chart ofFIG. 16.

FIGS. 18-19are plots illustrating gas exchange valve operation of a representative variable cam timing application using a device for selectively changing phase relationship between two or more rotating shafts with all cylinders activated according to the present disclosure. Line520represents piston position within a representative cylinder moving between top dead center (TDC) and bottom dead center (BDC).

The plot ofFIG. 18represents operation of a mechanical variable cam timing device according to the present disclosure with a baseline valve timing diagram similar to how a conventional camshaft operates. Line522represents the position or displacement of one or more exhaust valves per cylinder as they substantially simultaneously open and close relative to piston position line520. Line524represents the position or displacement of one or more intake valves as they substantially simultaneously open and close relative to piston position line520.

The plot ofFIG. 19illustrates how the duration of intake/exhaust valve opening events can be increased using a mechanical variable cam timing device in combination with a coaxial camshaft operating four valves per cylinder according to the present disclosure. Line530represents the position or displacement of a first exhaust valve while line532represents the position or displacement of a second exhaust valve on the same cylinder. Line534represents the position or displacement of a first intake valve and line536represents the position or displacement of a second intake valve on the same cylinder. As illustrated in the plot ofFIG. 19, a mechanical device to selectively change phase relationship between the crankshaft (as represented by the piston position) and the camshaft according to the present disclosure may be used to increase the overall valve opening times relative to the baseline timing as represented byFIG. 18.

InFIG. 20, an internal combustion engine610having a first bank612of cylinders and a second bank614of cylinders is shown. Only front pistons616and618are visible inFIG. 20. However, it is common for there to be 3 or 4 cylinders in each cylinder bank yielding V-6 or V-8 engines, respectively. Pistons616and618are coupled to a crankshaft624via connecting rods620and622. Pistons616and618reciprocate within cylinders626and628. The crank-slider mechanism of pistons616and618, connecting rods620and622and crankshaft624convert the linear motion of pistons616and618into rotary motion at crankshaft624. Above the pistons are combustion chambers630and631which are delimited by pistons616and618, cylinders626and628, and cylinder heads632and633. Within cylinder heads632and633, each cylinder is provided at least one intake valve and at least one exhaust valve. In the view shown inFIG. 20, only an intake poppet valve634is visible in cylinder head632and only an exhaust poppet valve636is visible in cylinder head633. Behind intake poppet valve634is an exhaust poppet valve; similarly, behind exhaust poppet valve636is an intake poppet valve. Exhaust poppet valve636is shown inFIG. 20as being depressed so that exhaust port640is open allowing fluid communication between combustion chamber631and exhaust manifold644. Similarly, intake poppet valve634is shown as being depressed so that intake port638is open allowing fluid communication between combustion chamber630and intake manifold646. Poppet valves636and638are normally in a non-depressed state, in which they are covering ports638and640, respectively. Valves634and636are biased toward the closed position by valve springs648and650, respectively. The spring forces are overcome when tappets652and654are depressed by cam lobes660and662coupled to camshafts656and658, respectively. When camshafts656and658are rotated such that cam lobes660and662are not in contact with tappets652and654, valves634and636are closed.

InFIG. 20, exhaust manifolds642and644are provided for cylinder heads632and633, respectively. However, as is typical, only one intake manifold646is provided with intake runners for all engine cylinders. The intakes from the two banks may remain separated for a distance upstream and combined further upstream.

Referring toFIG. 21, a single cylinder bank670having four reciprocating pistons674a-dis shown. Pistons674a-dreciprocate within cylinders (not shown. Each cylinder has an intake valve676a-dand an exhaust valve678a-dwhich are actuated by camshaft680having intake camshaft lobes682a-dand exhaust camshaft lobes684a-d, which press on tappets686a-hto depress valves676a-dand678a-d. Pistons674a-dare connected to crankshaft688via connecting rods. Crankshaft688is coupled to crankshaft sprocket690. Camshaft680is coupled to camshaft sprocket692. Crankshaft sprocket690drives camshaft sprocket692via belt drive694. A chain drive may be used in place of belt drive694. The drive ratio between camshaft sprocket692and crankshaft sprocket690is 2:1 so that camshaft680rotates at one-half speed of crankshaft688. In a V-8 configured engine, two banks of cylinders, such as shown inFIG. 21are coupled to the crankshaft and the crankshaft sprocket drives two camshafts.

Within camshaft680inFIG. 21is a coaxial, concentric camshaft (not specifically illustrated in this Figure) as previously illustrated and described with respect toFIGS. 1-19. Lobes682a-dare coupled to camshaft680with valves684a-dcoupled to the internal camshaft (such as a configuration shown inFIGS. 4-6), with the two camshafts being adjustable with respect to each other using any of a number of mechanical and/or hydraulic cam phasing devices, such as the mechanical device described herein. A camshaft adjuster (or phaser)693is shown coupled to camshaft680on the front of the engine. Adjuster693rotates both of the concentric camshafts when actuated. A camshaft adjuster695is coupled to the internal camshaft at the rear of the engine. Adjuster695rotates the internal camshaft with respect to camshaft680when actuated. In one embodiment, adjuster693is a hydraulic adjuster and adjuster695is a mechanical adjuster according to embodiments shown in the present disclosure. In another embodiment, both adjusters693and695are mechanical adjusters. An electronic control unit (ECU)696commands actuation of adjusters693and695based on current engine and/or ambient operating parameters and operating modes, such as a variable displacement operating mode, for example. In another alternative, lobes682a-dare coupled to the internal camshaft with lobes684a-dcoupled to camshaft680.

In yet another alternative, each cylinder has two intake valves and two exhaust valves actuated by two sets of coaxial camshafts such as shown inFIGS. 2A and 2B. One coaxial camshaft pair actuates all intake valves and the other coaxial camshaft pair actuates all exhaust valves.

Referring now toFIG. 22, valve and piston events are shown on a crank angle degree plot for a representative engine cycle with the cylinders activated (firing). The four strokes of the four-stroke engine are: expansion, exhaust, intake, and compression, each of approximately 180 degrees in duration. Thus, the engine completes one complete set of processes in 720 crank degrees, with the piston strokes occurring between a top center (TC) position and a bottom center (BC) position as represented by line700.

These strokes are used conventionally herein even when the cylinder is deactivated.

InFIG. 22, typical cam lifts are shown with the exhaust valve lift represented by line702and the intake valve lift represented by line704. During expansion, high pressure developed in the combustion chamber from the combustion of gases acts upon the piston to provide the power to drive the engine. Both valves are closed to contain the pressure, with the exhaust valve opening a bit before the piston reaches BC. A retreating piston along with both valves closed is shown inFIG. 22A. As the piston travels from BC to TC (FIG. 22B), the exhaust valve is open to expel burned gases into the engine exhaust. Just before the piston reaches TC, the intake valve is begins opening and both valves are partially open for some period of time, called valve overlap. During the intake stroke (FIG. 22C), the intake valve is open while the piston travels from TC to BC, thereby pulling in fresh air from the engine intake. Following intake, both valves are closed and the piston travels from BC to TC (FIG. 22D) to compress the gases in the combustion chamber in preparation for a combustion event.

According to an embodiment of the disclosure, overall efficiency of the engine is improved at modest torque demand when one bank of cylinders operates with normal valve timings, such as inFIG. 22, and the other bank of cylinders is effectively deactivated by commanding valve timings that are shown inFIG. 23. The exhaust valve timing is advanced as represented by line722inFIG. 23using a mechanical variable cam timing mechanism as previously described, relative to a typical valve event represented by line702inFIG. 22, such that the maximum lift position occurs at BC, i.e., the valve lift profile is substantially symmetrical with respect to BC. The intake valve lift, as represented by line724inFIG. 23, is retarded with respect to typical valve timing (line704inFIG. 22), such that the intake valve maximum lift occurs at BC and the lift profile is substantially symmetrical with respect to BC. The maximum lift of the exhaust valve occurs at the BC between expansion and exhaust and the maximum lift of the intake valve occurs at the BC between intake and compression.

Continuing withFIG. 23, during the first portion of expansion, the piston is traveling from TC to BC with both valves closed (FIG. 23A). In the middle portion of the expansion stroke, however, the exhaust valve begins to open. As the piston continues to move downward with the exhaust valve open, exhaust gases are pulled into the combustion chamber for the remainder of the expansion stroke,FIG. 23B. During the first part of the exhaust stroke, the piston passes through BC and begins upward movement toward TC with the exhaust valve open. The gases in the combustion chamber are expelled such as shown inFIG. 23C. The exhaust valve closes about the middle of the exhaust stroke. As both valves are closed from about 270 to 360 crank degrees, the gases trapped in the cylinder are compressed and a positive pressure develops in the cylinder. The valves remain closed for the first part of the intake stroke while the piston is traveling downwards from TC to BC, the expansion thereby relieving the pressure that developed during the later portion of the exhaust stroke. The later portion of the exhaust stroke and the early portion of the intake stroke are illustrated inFIG. 23D, showing both valves closed and the piston moving both upward during period D1and downward during period D2. During the later portion of the intake stroke, the intake valve lifts off its seat and starts to open. As the piston is moving toward BC, gases are pulled into the combustion chamber through the open intake valve from the engine intake,FIG. 23E. The intake remains open after BC and, when the piston moves upward from BC, gases are expelled from the combustion chamber into the engine intake,FIG. 23F. The intake valve closes in the middle portion of the compression stroke. With both valves closed, the gases in the combustion chamber are compressed during the later portion of the compression stroke,FIG. 23G. The pressure developed in the later portion of the compression stroke is relieved during the first portion of the next expansion stroke when the piston travels downward from TC.

Continuing to refer toFIG. 23, during portion B of the expansion stroke, a first quantity of gases are inducted from the exhaust into the combustion chamber. During portion C of the exhaust stroke a second quantity of exhaust gases are exhausted from the combustion chamber into the exhaust. In one embodiment of the present disclosure, valve timing is controlled using a mechanical camshaft adjustment device such that the first and second quantities are substantially equal. The timing at which the first quantity equals the second quantity coincides approximately with maximum lift occurring at BC between expansion and exhaust, or with the valve event approximately symmetrical with respect to BC. For the intake valve event, during portion E of the intake stroke, a third quantity of gases is inducted into the combustion chamber from the intake by virtue of the piston moving downward. During portion F of the compression stroke, a fourth quantity of gases is exhausted from the combustion chamber into the exhaust. By timing the intake event according to an embodiment of the present disclosure, the third and fourth quantities are equal such that there is no net flow from the intake into the combustion chamber. That timing is approximately symmetrical with BC between the intake and compression strokes or with maximum lift at that BC. As illustrated inFIG. 23, valve timing is controlled to provide cylinder deactivation with substantially no valve overlap.

In an alternative embodiment, valve timing is controlled to facilitate exhaust flow from the exhaust to the intake. Higher levels of exhaust gases can be tolerated without impairing combustion stability at relatively higher engine torque operating conditions. High torque load also corresponds to a mode of engine operation that produces a higher concentration of nitrous oxide gasses in the engine exhaust, making a higher flow of exhaust gas recirculation desirable. If exhaust gases flow into the intake of deactivated cylinders, such gases mix with intake gases for activated cylinders. Such backflow in deactivated cylinders is achieved by one or both of: advancing the exhaust camshaft timing slightly with respect to symmetrical timing shown inFIG. 23for the exhaust valve lift; and retarding the intake camshaft timing slightly with respect to the symmetrical timing shown inFIG. 23for the intake valve lift.

Cylinder pressure is represented by line730inFIG. 23. The pressure during periods B and C is substantially exhaust pressure and during periods E and F is substantially intake pressure. The pressure rises in the cylinders during periods D1and G, but is relieved in periods A and D2. As mentioned above, one problem in some prior art valve deactivation systems is that a substantial vacuum develops in the combustion chamber at periods in the cycle if both valves are caused to remain closed for multiple cycles. However, according to the disclosure, a significant vacuum does not develop over time because the valves open regularly to relieve the pressure, thereby mitigating the problem of oil pumping into the combustion chamber during the cylinder deactivation periods.

The valve events shown inFIG. 23are symmetrical with respect to BC. However, many engines are known to have slightly different cam profiles on the opening side and the closing side. In this case, the lifts cannot be exactly symmetrical. Also, it is known that there is inertia in causing the gases to flow. Thus, there is a slight delay between the valve opening and the gases starting to flow into or out of the combustion chamber. Similarly, once a flow is established, there is inertia in that flow. These flow dynamics become more important as engine speed increases. The desired effect, according to an embodiment of the present development, is to have no net flow of intake gases or exhaust gases through the engine. That is, the amount of air inducted from the intake during period E ofFIG. 23should be roughly equal to the amount of air expelled into the intake during period F. At some operating conditions, particularly at high engine speed, the valve timing to achieve no net flow is displaced a few degrees from being symmetrical about BC. Thus, according to an embodiment of the disclosure, the valve timing is adjusted to obtain substantially no net flow of intake or exhaust gases.

It is common to provide an external exhaust gas recirculation system including a duct between the engine exhaust and the engine intake. The duct has a valve to control the amount of exhaust gases inducted by the engine intake. According to another embodiment of the present disclosure, a small net flow of exhaust gases into the intake system can be provided by advancing the exhaust valve timing slightly and retarding the intake valve timing slightly, in relation to the valve timings shown inFIG. 23. This can be used as an additional supply to supplement the exhaust gas recirculation system flow or to supplant the exhaust gas recirculation flow for certain operating conditions.

In yet another embodiment, the exhaust valve timing is very slightly advanced and/or the intake valve timing is very slightly retarded (from what is shown inFIG. 23) so that the net flow is a slight exhaust flow into the intake. It is known that it is often undesirable to provide excess oxygen to exhaust systems in which a three-way catalyst is deployed. Selecting valve timings which cause a small exhaust flow to the intake ensures that no oxygen leaks to the three-way catalyst.

FIG. 24is a flowchart illustrating operation of a system or method for operating an engine in a variable displacement mode using variable cam timing to deactivate one or more cylinders according to the present disclosure. As those of ordinary skill in the art will understand, the functions represented by the flow chart blocks may be performed by software and/or hardware. Depending upon the particular processing strategy, such as event-driven, interrupt-driven, etc., the various functions may be performed in an order or sequence other than illustrated in the Figure. Similarly, one or more steps or functions may be repeatedly performed, or omitted, although not explicitly illustrated. In one embodiment, various functions are primarily implemented by software, instructions, or code stored in a computer readable storage medium and executed by a microprocessor-based computer or controller to control operation of the engine to selectively operate in a variable displacement mode by controlling a mechanical valve adjustment device to deactivate one or more cylinders. InFIG. 24, the engine is started at800and operation starts with all engine cylinders activated as represented by block802. In804it is determined whether engine conditions are favorable for operating in a variable displacement mode by deactivating one or more cylinders. In the illustrated embodiment, an entire cylinder bank is deactivated. Engine/vehicle and/or ambient operating conditions may include torque levels that are less than half of the maximum torque that the engine can produce, for example. In some applications, it may be undesirable to operate with only half of the engine cylinders firing due to vibration. Also, it may be desirable to wait until the engine warms up prior to deactivating one or more engine cylinders. If a negative results in804, control passes back to802. If a positive results in804, half (or one engine bank) of engine cylinders are deactivated by advancing exhaust valve timing so that the maximum lift is approximately at BC and retarding intake valve timing so that the maximum lift is approximately at BC to provide substantially no valve overlap.

As illustrated inFIGS. 1-24, a method for deactivating cylinders according to the present disclosure may include changing relative rotational phase between first and second rotating shafts in an internal combustion engine to provide substantially no valve overlap and no net flow of intake or exhaust gases during cylinder deactivation. In one embodiment, concentric intake and exhaust camshafts are phased to provide substantially no net flow of intake or exhaust gases during cylinder deactivation by biasing a first worm assembly102to apply a clockwise rotational bias torque between the first shaft110and second shaft112, and biasing a second worm assembly104to apply a counterclockwise rotational bias torque between the first and second shafts110,112to maintain the rotational phase between the first and second camshafts during steady-state rotation, i.e. other than during phase change actuation. An actuating torque is applied through the front side gear116and associated worm pinions120,122to the first and second worm assemblies102,104during phase change actuation to selectively change the rotational phase by advancing rotation of shaft112relative to shaft110while the shafts are rotating.

As such, the systems and methods of the present disclosure provide a mechanical variable camshaft timing device that can be used to adjust the phase relationship between two rotating shafts of an internal combustion engine to provide substantially no valve overlap and no net flow of intake or exhaust gases for deactivated cylinders. Various embodiments have the variable cam timing device adjusting the phase relationship between the camshaft and crankshaft and/or the phase relationship between coaxial camshafts operating intake and/or exhaust valves. Embodiments of the present disclosure provide compact packaging with desired reliability and durability such that the device can be implemented without increasing the length of camshaft bearings and with minimal or no overall increase in engine length. Opposite hand preload torque reduces or effectively eliminates backlash during operation of the device to reduce noise and wear. Friction locking within the advance/retard mechanism maintains the angular relationship between associated rotating shafts under dynamic loading during operation to reduce or eliminate need for ongoing adjustments by the actuating device. The virtually unlimited range of control of a variable cam timing device according to the present disclosure allows both the intake and the exhaust valves to be centered about different bottom centers such that there is no overlap to leak exhaust gasses into the intake manifold that is still operating firing cylinders on another engine bank. Furthermore, a variable cam timing device according to the present disclosure facilitates maintaining positive pressure within the deactivated cylinders to prevent oil from migrating past the piston rings into the combustion chamber.

While the best mode has been described in detail, those familiar with the art will recognize various alternative designs and embodiments within the scope of the following claims. While various embodiments may have been described as providing advantages or being preferred over other embodiments with respect to one or more desired characteristics, as one skilled in the art is aware, one or more characteristics may be compromised to achieve desired system attributes, which depend on the specific application and implementation. These attributes include, but are not limited to: cost, strength, durability, life cycle cost, marketability, appearance, packaging, size, serviceability, weight, manufacturability, ease of assembly, etc. The embodiments discussed herein that are described as less desirable than other embodiments or prior art implementations with respect to one or more characteristics are not outside the scope of the disclosure and may be desirable for particular applications.