Variable Valve Timing System

Abstract

A system for varying the valve timing of an internal combustion engine to increase the retarding effect of the engine for braking purposes, to increase the compression ratio of a relatively low compression ratio engine to enhance starting, and to optimize breathing to improve the engine performance over the useful operating speed range. The valve timing is varied by changing the total valve train length so as to shift the points on the cam profile at which the valve opening and closing events are determined.

Description

BACKGROUND OF THE INVENTION

This invention relates to variable valve timing systems and, more particularly, to systems for varying the valve timing of vehicle internal combustion engines to convert the engine into an air compressor to exert a braking effect on the drive train, to increase the effective compression ratio of a relatively low compression ratio engine to enhance starting, or to adjust the valve timing of a high speed engine to improve the performance throughout the operating speed range.

Various systems for varying valve timing have been proposed for some of the above and other purposes. Thus, the Lewis U. S. Pat. No. 754,466 shows an arrangement for relieving compression to enhance starting by manually extending an element mounted on the rocker arm to engage the exhaust cam and open the exhaust valve during a portion of the compression stroke.

Other devices for relieving compression to enhance starting are disclosed in the Jackson U. S. Pat. No. 1,172,362 and the Rounds U.S. Pat. No. 1,175,820. Here the exhaust cams are provided with an auxiliary relief or lobe which is circumferentialy spaced from the main lobe which opens the exhaust valve during the exhaust stroke. During normal operation the exhaust valve is not raised by the auxiliary lobe, but during starting the exhaust valve gear train is manually expanded so that the auxiliary lobe raises the exhaust valve during a portion of the compression stroke. In addition, the Rounds patent shows apparatus for manually adjusting the timing of the inlet and exhaust valves.

The Saurer U. S. Pat. No. 934,762 discloses an engine brake in which the exhaust cam is shifted circumferentially from its normal position to open during the "expansion" stroke, ignition being discontinued, so that air is compressed during the compression and "exhaust" strokes, and necessarily dumped at the beginning of the inlet and "expansion" strokes, so that the energy of the compressed air is not returned to the drive train during the expansion stroke.

The Kirchensteiner U.S. Pat. No. 1,637,118 and the Loeffler U.S. Pat. No. 1,947,996 disclose engine brakes in which the cam shaft is axially shifted for braking to de-activate the inlet valve and to drive the exhaust valve by a special double lobe cam, one lobe opening the exhaust valve during the intake stroke, while the other lobe dumps the compressed air near the end of the compression stroke. A graduated degree of braking is available in the Kirchensteiner engine brake by selectively inserting wedge elements beneath predetermined ones of the exhaust rocker arms to prevent the corresponding exhaust valves from closing, thereby eliminating the braking effect in the corresponding cylinders.

The engine brake according to the Ucko U.S. Pat. No. 2,002,196 obtains the results of the Loeffler brake without axially shifting the cam shaft. Rather, the rocker arm shaft is shifted eccentrically to render the push rods (and the inlet and exhaust cams) ineffective. An auxiliary double lobe exhaust cam is hydraulically coupled to the exhaust valve through a master piston, which is driven by the double lobe cam, and a slave piston which drives the exhaust valve rocker arm to open the exhaust valve during the intake and expansion strokes. A graduated braking effect is obtained by sequentially converting groups of one or more cylinders to air compressors.

The Cummins U.S. Pat. No. 3,220,392 discloses another engine braking system employing hydraulically coupled master and slave pistons, the slave piston driving the exhaust valve rocker arm, and the master piston being driven by an auxiliary exhaust cam, the injector rocker arm of the corresponding cylinder, or by the inlet or exhaust rocker arm of another cylinder, so as to dump compressed air at or near the end of the compression stroke. Unlike Ucko, however, the Cummins mechanism for opening an exhaust valve at or near top dead center does not interfere with the actuation of the exhaust valve by the normal exhaust valve actuating mechanism. Nevertheless, the independent mechanism for actuating the exhaust valve for braking requires considerable additional structure, thus increasing the complexity and cost of that engine brake. Furthermore, hydraulic coupling between the exhaust rocker arm of one cylinder and the inlet or exhaust rocker arm of the appropriate other cylinder would be difficult to arrange with a V-8 engine.

In the engine brake according to the Jones et al. U.S. Pat. No. 3,439,662 a single auxiliary cam sequentially drives the master pistons, which in turn actuate the corresponding slave pistons to open the exhaust valves at the end of the compression stroke. Apparatus is included to change the timing of the opening of the exhaust valves in accordance with the engine speed in order to increase the braking effect with increasing engine speed.

The Siegler U.S. Pat. No. 3,547,087 discloses another engine brake employing a mechanism external to the intake and exhaust valve gear train, but in this system a solenoid operated hydraulic valve remote from the engine brake mechanism is actuated to pump up a piston so as to block the return movement of the rocker arm, thereby holding the intake or exhaust valve partially open throughout the braking period.

The Haviland U.S. Pat. No. 3,332,405 shows an engine brake in which the exhaust valve is opened at the end of the compression stroke by a separate engine braking cam when a plunger mounted in the rocker arm is hydraulically pumped up to engage the braking cam in response to a remote solenoid valve. In an effort to improve the response time of the system, a separate low pressure oil supply is required to keep the lines filled with oil.

The Jonsson U.S. Pat. No. 3,367,312 discloses an engine braking system in which the normal base circle of the exhaust cam is relieved to form an auxiliary base circle, the transition between the two base circles constituting an auxiliary ramp displaced circumferentially from the normal opening ramp, so that when the lash is removed from the exhaust valve train, the exhaust valve is opened by the auxiliary ramp at the end of the compression stroke. The lash is removed by a plunger mounted in the rocker arm which may be hydraulically extended when a remote valve is manually actuated to communicate the plunger with the lubrication pump. Inasmuch as there is no mechanism for hydraulically locking the plunger in the extended position, the rotating exhaust cam will reciprocate the plunger in its cylinder despite the hydraulic force supplied by the lubrication pump, thereby substantially impairing the performance of the engine brake. Furthermore, a very large force is applied to the exhaust valve and the plunger when the piston travels through its compression stroke, such force being a function of speed, exhaust valve opening, exhaust valve diameter and compression ratio. In a diesel engine such force would greatly exceed the opposite force on the plunger developed by the engine lubricating pump, so that the plunger would be collapsed and the desired braking effect minimized.

The Muir U.S. Pat. No. 3,525,317 discloses an engine brake providing a graduated braking effect by arranging a multiple-position switch for operation as the throttle pedal is retracted beyond the idling position. At the first position the fuel is cut off to create "motoring" friction, at the second position the exhaust valves are held continuously in a partially open position, and at the third position a butterfly valve in the exhaust mainfold is actuated to provide back pressure therein.

The Sweat U.S. Pat. No. 2,806,459 discloses an intricate device for changing the timing of motor valves in accordance with the speed of the motor by adjusting the position of the rocker arm fulcrum and thereby adjusting the clearance in the valve train and the amount of valve opening. The rocker arm fulcrum is driven by a motor, the electrical contacts for which are operated by a piston displaced by air pressure generated by a fan driven by the motor. The cam shaft is arranged to provide advanced timing when the clearance in the valve train is small, at high speeds, while at lower speeds the clearance is larger and the valve timing is thus retarded.

The Lieberherr U.S. Pat. No. 2,936,575 discloses apparatus for varying the valve timing of a supercharged gas engine in accordance with the pressure of the intake manifold or the governor fuel control shaft in an effort to obtain an approximately constant air-fuel ratio at all loads. The timing is varied by lateral displacement of the cam follower in response to the intake manifold pressure or the fuel control shaft.

The Ostborg U.S. Pat. No. 3,224,423 shows a valve timing system in which the timing of the inlet and exhaust valves is varied in accordance with the intake manifold pressure, the phase of the inlet and exhaust camshafts being shifted in opposite directions with respect to that of the crankshaft by means of planetary gear systems.

The above conventional variable valve timing devices are relatively complex in construction, expensive to manufacture and install, susceptible to malfunction, and require relatively frequent servicing. In the prior art engine brakes using hydraulic circuitry, the control valve is remote from the operating pistons, providing an inherently slow reaction time. This can be disastrous to the system, because if the engine is allowed to fire while the engine brake is on, the cylinder pressure will destroy the valve train. The time required to deactivate the engine brake is therefore extremely important.

Another difficulty with conventional engine brakes results from the fact that the fuel condition is not sensed before the brake is activated. Rather, the fuel is shut off and the engine brake activated through separate control circuitry, so that the valve train will be destroyed if the fuel shut-off circuitry malfunctions.

SUMMARY OF THE INVENTION

It is an object of the present invention to overcome the above-mentioned difficulties and shortcomings of conventional systems for varying the valve timing of an internal combustion engine. Another object of the invention is to provide novel methods and apparatus for selectively controlling the timing events of one or more valves of an internal combustion engine. A further object of the invention is to provide a reliable, but relatively simple, system for converting an internal combustion engine into an air compressor to enhance engine retardation for braking purposes. A still further object of the invention is to provide a vehicle engine braking system which has a rapid response time and which may be readily incorporated in any internal combustion engine with minimum modification thereof. Yet another object of the invention is to provide an engine braking control system which ensures that the engine brake is deactivated while fuel is being delivered to the engine. Still another object of the invention is to provide a variable valve timing mechanism to provide additional breathing at the higher engine speeds so as to improve the performance throughout the operating speed range. Still a further object of the invention is to provide a variable valve timing system which enhances starting of internal combustion engines of relatively low compression ratio by increasing the effective compression ratio.

These and other objects of the invention are attained by selectively changing the total valve train length so as to shift the points on the cam profile at which the valve opening and closing events are determined. The change in total valve train length is effected by a novel hydraulic lash adjuster which forms a portion of the valve train. The hydraulic lash adjuster is made to expand and lock hydraulically, or collapse, upon command, thereby regulating the lash in the valve train.

The lash adjuster includes a piston which is spring biased in a direction to expand the lash adjuster and thereby lengthen the valve train, the piston being hydraulically locked in the expanded position when a ball check valve is permitted to seat and seal a high pressure chamber to which oil under pressure is admitted past the check valve. The oil under pressure is also constantly applied to one side of a control piston, on the opposite side of which is exerted a biasing force and a variable hydraulic pressure which may be selectively varied between atmospheric pressure and the same pressure as applied to the first side of the control piston. When the variable hydraulic pressure is at atmospheric pressure, the force on the first side of the control piston overcomes the biasing force and displaces the control piston away from the ball check valve, which may then seat to lock the lash adjuster in its expanded state.

On the other hand, when the variable hydraulic pressure is raised so that the pressures on both sides of the control piston are equal, the biasing force displaces the control piston to unseat the ball check valve, thereby collapsing the hydraulic lash adjuster.

The hydraulic lash adjuster is preferably mounted in the rocker arm coupled to the valve whose timing is to be varied. The rocker arm shaft is provided with two longitudinal passageways, one of which is communicated with a source of oil under pressure. This passageway supplies oil under pressure past the ball check valve into the high pressure chamber of each of the various hydraulic lash adjusters, and also supplies oil under pressure to one side of the control piston thereof. A solenoid operated valve selectively communicates the first passageway in the rocker arm shaft with the second passageway therein, which in turn pressurizes the oil on the opposite side of the control piston in each lash adjuster when permitted to do so by the solenoid actuated valve. A very rapid reaction time is attained for the hydraulic lash adjusters by mounting the control valve in the rocker arm shaft.

When the variable valve timing system is sued for engine braking, the hydraulic lash adjusters are mounted in the rocker arms for the exhaust valves, and the normal base circle of each exhaust cam is relieved to provide a secondary cam base circle, thereby forming secondary opening and closing profiles between the normal and secondary base circles. The secondary opening profile is circumferentially located so as to open the exhaust valve at or near the end of the compression stroke when the hydraulic lash adjuster is expanded to substantially eliminate the valve lash.

It is absolutely necessary that the engine brake not be operated while fuel is being delivered to the engine, since ignition could occur in a compression-ignition engine. This requirement is satisfied by including in the electrical control circuitry which energizes the solenoid operated valve mounted on the rocker arm shaft a switch which is closed when the injection pump rack is in its dead rack or no fuel position. Thus, the engine brake can only be energized when the vehicle overruns the particular setting of the throttle, in which condition the governor moves the injection pump rack to the dead rack position. As the engine decreases to idle condition, the governor then moves the rack to turn on idle fuel, but before the fuel is turned on, the rack opens the switch to deactivate the engine brake. Accordingly, the engine can never stall, because the engine brake will not function below idle speed.

The variable valve timing system is employed to enhance the starting of a relatively low compression ratio engine, by mounting the hydraulic lash adjusters in the rocker arms which actuate the inlet valves, so that the total length of the inlet valve train may be selectively extended or contracted. This displaces the locations along the profile of each inlet cam at which the opening and closing events of the corresponding inlet valve are determined. With the last adjusters extended so as to lengthen the total inlet valve trains, the inlet valves close in the normal manner somewhat after bottom dead center, at which time the compression actually begins. Thus, the effective compression ratio is somewhat lower than the theoretical compression ratio which would be attained if the inlet valve closed at bottom dead center. To enhance the starting of the relatively low compression ratio engine, each hydraulic lash adjusterm is collapsed to provide an excess inlet valve clearance sufficient to displace the closing point of the inlet valve to bottom dead center, so that the actual compression ratio under such starting conditions equals the theoretical compression ratio.

In order to prevent any possible damage to the inlet valve gear components when the engine operates with excess inlet valve clearances, a throttle limiting member is automatically extended to limit the maximum engine speed when the hydraulic lash adjusters are collapsed. After the engine has started and warmed up, the hydraulic lash adjusters are expanded to provide normal inlet valve clearance, and simulataneously the throttle limiting member is retracted to permit normal full throttle operation.

The variable valve timing system can also be employed to increase the performance of internal combustion engines throughout the operating speed range by increasing the valve lift of the inlet and/or exhaust valves, as well as the time period during which they are open, at high engine speeds to provide additional breathing. In this application, the normal cam base circles of the corresponding valves are relieved to provide auxiliary opening and closing profiles which are mutually spaced apart to a greater extent than the opening and closing profiles which are effective in the lower speed range. Thus, when the hydraulic lash adjusters are expanded for the higher engine speeds, the opening and closing of the corresponding valves is determined by the auxiliary opening and closing profiles, thereby increasing the valve lift and the period during which such valves are open.

BRIEF DESCRIPTION OF THE DRAWINGS

All of the above is more fully explained in the detailed description of the preferred forms of the invention which follows, this description being illustrated by the accompanying figures of the drawings, in which:

FIG. 1 is a chart illustrating the pressure-volume relationship for a four cycle internal combustion engine during "motoring," when ignition does not occur;

FIG. 2 is a chart illustrating the change in the pressure-volume relationship when the cylinders are converted to compressors to effect engine braking;

FIG. 3 is a chart illustrating cylinder pressure and inlet and exhaust valve lift vs. crank angle under normal operation, when ignition is prevented to provide motoring friction, and during engine braking;

FIGS. 4A, 4B, and 4C illustrate the cam profiles of the inlet and exhaust cams in the variable valve timing mechanism according to the present invention for engine braking;

FIG. 5 is a schematic illustration, partially in vertical section, of the variable valve timing system employing cams formed in accordance with FIGS. 4A, 4B, and 4C;

FIG. 6 is a chart illustrating the brake horsepower and torque developed under normal operation, as well as the negative horsepower obtained from motoring friction and from the engine braking system of FIG. 5;

FIG. 7 is a view taken along the line 7--7 of FIG. 5 and looking in the direction of the arrows;

FIG. 8 is a partial side elevational view, partly in section, of the rocker arm shaft of FIG. 7;

FIGS. 9A and 9B are views taken along the line 9--9 of FIG. 7 and looking in the direction of the arrows, showing the hydraulic lash adjuster in contracted and expanded states, respectively;

FIG. 10 is an enlarged sectional view of the hydraulic lash adjuster of FIG. 9;

FIGS. 11A, 11B, 11C and 11D are schematic illustrations of alternative electrical control circuitry suitable for employment with the engine braking system of FIGS. 4-10;

FIG. 12 is a schematic illustration, partially in vertical section, of another embodiment of the variable valve timing system for engine braking;

FIG. 12A is an enlarged view, in vertical section, of the hydraulic lash adjuster shown in FIG. 12;

FIG. 13 is a schematic diagram illustrating the opening durations of the inlet and exhaust valves during the normal operation of an internal combustion engine;

FIG. 14 is a chart illustrating the inlet valve timing when excess clearance is provided by the variable valve timing system according to another embodiment of the invention;

FIG. 15 is a schematic diagram illustrating the inlet valve opening duration with the excess clearance shown in FIG. 14;

FIG. 16 is a chart illustrating the pressure characteristics in the cylinders during the compression and expansion strokes with normal and excess inlet valve clearances;

FIG. 17 is a schematic illustration of the relative position of the piston in the bore when the inlet valve closes under normal conditions;

FIG. 18 is a schematic illustration of the relative position of the piston in the bore at the time that the inlet valve closes when excess clearance is provided;

FIG. 19 is a schematic illustration, partially in vertical section, of the variable valve timing system for producing the effects illustrated in FIGS. 14, 15, 16 and 18, thereby increasing the compression ratio in order to enhance the starting of a relatively low compression ratio engine;

FIG. 19A is a partial side elevational view, partly in section, of the rocker arm shaft of FIG. 19;

FIG. 20 is a schematic illustration of a variable valve timing system for engine braking and for enhancing the starting of a relatively low compression ratio engine;

FIG. 21 is another system employing the variable valve timing mechanism for engine braking and for enhancing the starting of a relatively low compression ratio engine;

FIGS. 22A, 22B and 22C illustrate the cam profiles of the inlet and exhaust cams employed in another embodiment of the variable valve timing system for improving the performance of internal combustion engines throughout the operating speed range;

FIG. 23 is a schematic illustration, partially in vertical section, of the variable valve timing system employing cams formed in accordance with FIGS. 22A, 22B and 22C;

FIG. 23A is a partial side elevational view, partly in section, of the rocker arm shaft of FIG. 23;

FIG. 24 is a chart illustrating the volumetric efficiency attained with the system of FIG. 23; and

FIG. 25 is a schematic illustration of a variable valve timing system for engine braking and for increasing the performance of an internal combustion engine throughout the operating speed range.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

As the gross weights and road speeds of vehicles increase, greater demands are placed upon the braking systems. It is advantageous to use the friction of the drive train and the engine for its retardation effect to aid in vehicle control, espcially under downhill or coasting conditions, in order to reduce dependence on the conventional service brakes, prolong the useful life of the brake components, and promote safety.

A light braking effect is obtained from an internal combustion engine by preventing ignition, either by deactivating the ignition circuit in a spark-ignition engine, or by stopping the injection of fuel in a compression-igniton engine. During such "motoring," the vehicle drives the engine, and retardation occurs because of the friction of the moving parts and because of the pneumatic forces on each piston as it moves in its cylinder.

In FIG. 1, the curve 31 represents the pressure-volume relationship during motoring, and it is apparent that the compression and expansion portions of the curve are very similar. Accordingly, the negative work, or power absorbed by the gases, during compression only slightly exceeds the positive work exerted by the gases on the pistons during expansion. Thus, most of the negative work occurs during the intake and exhaust strokes when air is drawn in, or forced out, through the constrictions surrounding the open inlet and exhaust valves, respectively.

The positive work developed during the expansion stroke may be minimized to enhance engine retardation by converting the engine into an air compressor by opening the exhaust valve at or near the completion of the compression stroke to dissipate the energy of the compressed air to atmosphere. Such compression release type of engine braking provides the pressure-volume characteristic shown by curve 32 in FIG. 2. Here it may be seen that negative work still occurs during the compression stroke until the exhaust valve is opened to allow the gases to expand into the exhaust manifold. The pressure then drops rapidly, so that the gases no longer do positive work on the piston during the expansion stroke. If the engine is equipped with a turbocharger, the "dumped" gases drive the turbocharger and thus increase the charging pressure, so that the pressure during the intake stroke is somewhat higher during engine braking than it is during motoring. The resulting increased air flow through the engine, drawn in during the intake stroke and exhausted during the exhaust stroke, results in additional negative work or power absorbed.

The upper portion of the chart shown in FIG. 3 shows cylinder pressure characteristics as a function of the position of the crankshaft, curve 34 being the pressure characteristic during ignition, curve 35 the pressure characteristic during "motoring," when there is no ignition, and curve 36 showing the pressure characteristic when the exhaust valve is opened near the end of the compression stroke to dump the compressed air and effect engine braking.

The lower portion of the chart of FIG. 3 shows the valve lift of the inlet and exhaust valves of an internal combustion engine provided with the variable valve timing system according to the present invention for opening the exhaust valve near the end of the compression stroke to effect engine braking. Curve 37 shows that the inlet valve timing under all conditions remains the same as that found in any four stroke cycle engine. There are two different timings for the exhaust valves, however, depending on whether or not the engine brake is activated. Thus, curve 38 shows the exhaust valve lift when the engine brake is off, while curve 39 shows that when the engine brake is on, the exhaust valve opens near the end of the compression stroke shortly before top dead center.

The valve lifts illustrated in FIG. 3 are obtained from the cam profiles shown in FIGS. 4A, 4B and 4C. The conventional inlet valve lift 37 is obtained from the conventional inlet cam 40 (see FIG. 4B), while the profile of the exhaust cam 42 (see FIG. 4C) is relieved at 43 to provide a secondary base circle. Accordingly, there is provided between the secondary base circle 43 and the conventional base circle 44 an auxiliary opening profile 46 and an auxiliary closing profile 47.

When the exhuast exhaust gear train is expanded so as to provide only the lash 48 to accommodate changes in valve train length due to temperature variations, the exhaust valve will open slightly before top dead center where indicated on the auxiliary opening profile 46, and will close where indicated on auxiliary closing profile 47, to provide the valve motion shown by curve 39. In addition to extending the time during which the exhaust valve is open during the engine braking mode, the reduction in lash clearance resulting from the extension of the exhaust valve train causes the additional exhaust valve lift indicated at 50.

When additional valve lash is introduced into the exhaust valve gear train, so as to provide the total lash indicated at 52, the auxiliary opening and closing profiles 46 and 47 are rendered ineffective. The normal exhaust valve timing is then obtained, providing the valve lift shown by the curve 38 in FIG. 3.

The variable valve timing system employing the exhaust cam 42 for engine braking is shown in FIG. 5. The cam 42 is formed on a camshaft 55 which is conventionally driven by a crankshaft 56, which in turn is driven by eight pistons 58 through corresponding connecting rods 59 in a V-8 engine in the illustrated embodiment, only one piston and associated elements being shown for simplicity. Each piston 58 reciprocates in a conventional sleeve 60, which is mounted in the engine block 61, to which are secured four cylinder heads 63, 64, 65, and 66, which mount rocker arm shafts 67, 68, 69, and 70, respectively. Each rocker arm shaft carries the inlet and exhaust rocker arms for two cylinders.

Mounted for reciprocation in the cylinder head 63 is an exhaust valve 72, which is driven by one end of an exhaust rocker arm 74, which is mounted for rotation on the rocker arm shaft 67. A valve seat insert 73 is suitably recessed so that the exhaust valve will not engage the piston when the valve train is expanded for engine braking, during which time the exhaust valve is provided with additional lift as shown by curve 39 of FIG. 3.

The opposite end of the rocker arm 74 mounts a hydraulic lash adjuster 76, which will be described in detail hereinafter. The lash adjuster 76 receives the upper end of a pushrod 78, the lower end of which is received in a cam follower 80, which engages the exhaust cam 42. The pushrod is preferably of somewhat heavier construction than those used to rock the inlet rocker arms, because of the forces exerted on the exhaust valve train when the exhaust valve is opened near the end of the compression stroke.

The hydraulic lash adjuster 76 is threadedly received in the rocker arm 74, so that the lash adjuster may be positioned in the rocker arm when the exhaust valve train components are cold so as to provide sufficient lash to accommodate the thermal expansion of the valve gear train elements. The last adjuster is locked in the desired position by a lock nut 82.

A conventional solenoid 85 mounted on the rocker arm shaft 67 may be energized to expand the hydraulic lash adjusters 76 mounted in the exhaust rocker arms 74 and 74a, as will be explained in detail hereinafter, so that the exhaust valves 72 of the two cylinders associated with the cylinder head 63 are opened by the auxiliary profiles 46 of the exhaust cams 42 to dump the compressed air in these cylinders near the end of the compression stroke. When the solenoid 85 is de-energized, the associated hydraulic lash adjusters 76 are collapsed, and the corresponding exhaust valves open near the end of the expansion stroke, thereby discontinuing the engine braking. Similarly, solenoids 86, 87 and 88, identical in construction and operation to the solenoid 85, are mounted on the rocker arm shafts 68, 69 and 70, respectively, and may be selectively energized or de-energized to expand or collapse, respectively, the hydraulic lash adjusters 76 mounted on the corresponding rocker arms 74 and 74a to shift the timing of the exhaust valves of the corresponding cylinders.

The solenoids 85, 86, 87 and 88 may be selectively energized by the switches 93, 94, 95 and 96, respectively, which are mounted so as to be sequentially actuated by the brake pedal 98, provided that a manual cab switch 100 and a limit switch 102 are closed. The cab switch 100 and the limit switch 102 are series connected between each of the switches 93-96 and a suitable source 104 of electric power, such as the conventional vehicle battery. The cap switch 100 enables the operator to override or de-activate the automatic operation of the engine brake.

The limit switch 102 is mounted in operative relation to the control rack 106 of a conventional fuel injection pump manufactured by American Bosch Arma Corporation, which is provided with a fuel limiting governor identified as the GVB/C governor, manufactured by American Bosch Arma Corporation and conventionally supplied with the pump. Briefly, the governor includes a plurality of flyweights 110 driven by the crankshaft 56, and as the weights revolve, centrifugal force tends to move them outwardly, this movement being opposed by a governor spring 112 acting through a sliding sleeve assembly 114. The sliding sleeve assembly provides the pivot for a fulcrum lever 115, one end of which is positioned by a control or throttle lever (not shown), and the other end of which positions the control rack 106. Displacement of the control rack rotates the fuel injection plungers (not shown), the rotational disposition of which determines the quantity of fuel delivered with each stroke of the pumping mechanism through a plurality of high pressure fuel lines 117 to a plurality of injector nozzles 118, one of which is provided for each combustion chamber of the engine, as is well known to those skilled in the art.

The limit switch 102 is located so that it is only closed when the control rack 106 is in its dead rack or zero fuel delivery position. Accordingly, the engine brake control system senses the delivery of fuel to the combustion chambers, and ensures that the engine brake cannot be activated when fuel is being delivered to the combustion chambers. This not only prevents wasted fuel during engine braking, but more importantly prevents the damage to the valve train which would result if combustion occurred while the exhaust valve is open, such combustion being possible in a compression-ignition engine even though the exhaust valve is open.

The switches 93-96 are so disposed with respect to the brake pedal 98 that they are sequentially closed before the vehicle service air brake is activated. The chart shown in FIG. 6 illustrates the graduated or modulated braking effect available with the system shown in FIG. 5. Thus, the curves 120 and 121 show the brake horespower and torque, respectively, developed by a representative V-8 diesel engine under normal operation.

If the operator wishes to retard the vehicle, either during downhill operation or when approaching a traffic light or intersection on level terrain, for example, he lifts his foot from the throttle pedal (not shown) and the governor will drive the control rack 106 to the dead rack or no fuel position. This terminates combustion and effects engine or motoring friction, during which condition the negative or absorbed horespower is proportional to the engine RPM, as illustrated by the curve 122. When the rack is in the no fuel position, the limit switch 102 is closed, permitting activation of the engine brake providing that the manual override switch 100 is closed.

Should the operator desire additional braking, he depresses the brake pedal 98 to close the switch 93 thereby converting the two cylinders associated with the cylinder head 63 to air compressors and thus obtaining the retarding effect shown by the curve 123. If the brake pedal 98 is further depressed to close the switch 94 (while maintaining the switch 93 in its closed position), the solenoid 86 is energized to convert the two cylinders associated with the cylinder head 64 to air compressors, thus obtaining the retarding effect shown by the curve 124. Similarly, when the brake pedal is further depressed to close the switches 95 and 96 (the switches 93 and 94 remaining closed), the solenoids 87 and 88 are energized and an engine braking effect is obtained from the cylinders associated with the cylinder heads 65 and 66, respectively, thus obtaining the braking effects illustrated by the curves 125 and 126, respectively.

Should still additional braking be desired, the brake pedal 98 is further depressed to actuate an air valve 128 and thus activate the conventional service brake (not shown). Thus, the present engine braking system provides the operator with great flexibility to obtain the desired degree of braking in accordance with the particular downhill or other condition.

The rocker arm shaft assembly for the rocker arm shaft 67 is shown in FIGS. 7 and 8, it being understood that the corresponding assemblies for the other rocker arm shafts are identical. The rocker arm shaft 67 is supported on the cylinder head 63 by a pair of spaced brackets 130 and 131. A pair of inlet rocker arms 133 and 133a are mounted for rocking motion on the rocker arm shaft 67 and are urged against the brackets 130 and 131 by a pair of spring elements 135, anchoring portions of which are received in a pair of grooves 137 and 138, respectively, formed in the rocker arm shaft 67.

Similarly, the exhaust rocker arms 74 and 74a are urged against the brackets 130 and 131 by a pair of springs 140, the remote ends of which bear against a pair of washers 141 which are retained by a pair of snap rings 142, the snap rings being received in a pair of grooves 143 and 144, respectively, formed in the rocker arm shaft 67.

Each of the inlet rocker arms 133 and 133a is driven by a pushrod 146 (shown in phantom), the upper end of which is received in an adjusting screw 148 which is threadedly received in the inlet rocker arm. The position of the adjusting screw 148 is set when the inlet valve train components are cold so as to provide sufficient lash to accommodate the expansion of the valve train elements when their temperature rises. The adjusting screws are locked in the desired position by a locking nut 150. The lower end of each pushrod 146 is received in a cam follower (not shown) which rides upon a corresponding inlet cam (not shown) formed on the camshaft 55.

The rocker arm shaft 67 is formed with two longitudinal passageways 154 and 155, which are plugged at each end by a plug 156 and 157, respectively. Oil under pressure is supplied through a conduit 158 to the passageway 155 from a conventional oil pump 160 (see FIG. 5), which draws oil from the engine sump 161 through a conduit 162, the conduit 158 communicating with the passageway 155 through a bore 164 in the bracket 131. The oil pump 160 may also supply oil to bearings and other engine components (not shown) for purposes of lubrication and/or cooling. The bore 164 includes a threaded reduced diameter portion 165 threadedly receives a plug 166, which extends into and is snugly received in a lateral bore 168 formed in the rocker arm shaft 167 and communicating with the passageway 155. The plug 166 locks the rocker arm shaft 67 against rotational or axial displacement with respect to the brackets 130 and 131, and a bore 170 extending through the plug 166 communicates the conduit 158 with the oil gallery 155.

Oil under pressure flows from the oil gallery 155 through a pair of lateral bores 173 and 174 to lubricate the rocking motion of the inlet rocker arms 133 and 133a on the rocker arm at 67, and also shows through a pair of lateral bores 175 and 176 to lubricate the rocking motion of the exhaust rocker arms 74 and 74a and to supply oil under relatively constant pressure through a lower bore 178 in the exhaust rocker arms to the hydraulic lash adjusters 76 (see FIG. 9).

The exhaust rocker arms 74 and 74a are grooved (see groove 180 in the exhaust rocker arm 74a as viewed in FIG. 9) so as to permit sufficient oil to pass through a pair of bores 181 and 182, respectively, to drip over the noses of the exhaust rocker arms and thus lubricate the slipper faces 183 and 184 thereof. Similarly, the intake rocker arms 133 and 133a are grooved to supply oil through a pair of bores 185 and 186, respectively, therein so as to lubricate the slipper faces thereof.

The rocker arm shaft 67 is formed with a lateral bore 189 to threadedly receive a control valve 190 by means of which oil under pressure may be selectively communicated from the oil gallery 155 to the "on-off" oil gallery 154. The control valve 190 includes a valve body 191 upon which is mounted the solenoid 85, the valve body being formed with a central bore 192 into the which the solenoid armature 193 extends. Mounted in the upper portion of the bore 192 is an upper valve seat 194 having a central bore 195 so formed that an annular space exists between the bore 195 and the armature 193 for all positions of the armature. A lower valve seat 197 is mounted in the lower portion of the bore 192, and a ball valve element 198 is disposed between the upper valve seat 194 and the lower valve seat 197.

A plurality of lateral bores 200 in the valve body communicate the portion of the bore 192 between the upper and the lower valve seats with the "on-off" oil gallery 154. The lower portion of the bore 192 communicates with the "constant" oil supply gallery 155, a seal 202 ensuring that oil may only flow between the two galleries 154 and 155 through the central bore 192 and the lateral bore 200. Thus, when the ball valve 198 is displaced from the lower valve seat 197, communication is established between the oil galleries 154 and 155. A plurality of upper lateral bores 204 in the valve body 191 communicate the upper portion of the bore 192 above the upper valve seat 194 with atmosphere, so that when the ball valve 198 is displaced from the upper valve seat 194, the "on-off" oil gallery 154 is exposed to atmospheric pressure.

When the solenoid 85 is de-energized, the pressurized oil in the gallery 155 (at the "constant" pressure established by the oil pump 160) drives the ball 198 upwardly against the upper valve seat 194, simultaneously displacing the armature 193 upwardly. (Alternatively, the upper valve seat 194 could support a light spring for biasing the armature 193 upwardly.) Pressurized oil then flows through the bore 192 and the lateral bores 200 to the "on-off" gallery 154, and through a pair of lateral bores 206 and 208 in the rocker arm shaft 67 to a pair of upper bores in the rocker arms 74 and 74a, FIG. 9 showing the upper bore 210 in the rocker arm 74a.

The upper bore 210 communicates with the hydraulic lash adjuster 76, and when the pressure of the oil in the bore 210 is raised to that of the pressurized oil in the constant supply gallery 155 by communicating with the bore 210 with the gallery 155 (see FIG. 9A), the hydraulic lash adjuster 76 is in its collapsed state, as will be explained hereinafter.

When the solenoid 85 is energized, the armature 193 drives the ball 198 against the lower valve seat 197, sealing the "on-off" gallery 154 from the constant oil supply gallery 155. At the same time, the oil in the gallery 154 is exposed to atmospheric pressure through the lateral bores 200, the bore 195 and the lateral bores 204 (see FIG. 9B). When the upper bores 210 are vented to atmospheric pressure, the hydraulic lash adjusters 76 expand, as will be explained hereinafter.

Referring now to FIGS. 9 and 10, the hydraulic lash adjuster 76 includes a body 215, the outer lateral surface of which is threaded so as to permit vertical adjustment in a complementary bore 216 formed in the exhaust rocker arms 74 and 74a, in order to obtain the desired lash to accommodate for the thermal expansion of the valve gear train components, as mentioned above. The lateral surface of the body 215 is also formed with a pair of spaced annular grooves 217 and 218, which communicate with the bores 178 and 210, respectively, in the exhaust rocker arms. A pair of lateral bores 219 and 220 in the body 215 communicate the annular groove 217 with a central chamber 222, and a lateral bore 224 communicates the groove 218 with an upper chamber 226, a control piston 228 being disposed between the central chamber 222 and the upper chamber 226.

Oil is prevented from escaping from the upper chamber 226 by a plug 229 which is retained by a snap ring 230, the snap ring being seated in a complementary groove formed in the body 215. A seal ring 231 provides a liquid-tight seal between the plug 229 and the body 215. A compression spring 223 acts between the plug 229 and the control piston 228 to bias the latter downwardly, the control piston being bored at 235 to receive the compression spring. The control piston 228 is reciprocable between the plug 229 and stop 237 formed in the interior of the body 215.

A finger 238, depending upon the control piston 228, extends into a port 239 communicating the central chamber 222 with a high pressure chamber 240. A perforated ring carrier 242 disposed in the high pressure chamber supports a light compression spring 243 which biases a ball check valve 245 upwardly to seal the port 239.

The high pressure chamber 240 is closed by a socket piston 247, which is slidably received in a cylinder 248 formed in the lower portion of the body 215. The socket piston 247 is formed with a socket 249 in the lower portion thereof adapted to receive the upper end of the pushrod 78, and the socket piston is biased downwardly (as viewed in FIGS. 9 and 10) by a compression spring 250. The compression spring 250 is received in a pocket 251 formed in the upper portion of the socket piston, the upper end of the compression spring 250 being received in an outwardly extending flange 252 of the spring carrier 242. The compression spring 250 is relatively stiff compared with the ball valve biasing spring 243, and so retains the spring carrier 242 against the lower face 254 of the inwardly extending portion of the body 215 which forms the port 239.

The socket piston 247 may reciprocate between a snap ring 256, which is seated in a complementary groove formed in the lower interior portion of the body 215, and a stop 258 formed in the body 215 at the upper end of the cylinder 248. The outer lateral surface of the socket piston 247 includes a plurality of spaced annular grooves 259, which retain oil therein to lubricate the movement of the socket piston and thus extend its life. The bleed down rate of the oil in the high pressure chamber 240 past the socket piston when it is hydraulically locked is sufficiently low so that movement of the socket piston is negligible during the valve event. Furthermore, as soon as the cam follower 80 engags the secondary base circle 43, the socket piston is restored to its fully expanded position.

The hydraulic lash adjuster 76 is made to expand and collapse as follows. when the solenoid 84 is de-energized, pressurized oil from the constant oil supply gallery 155 in the rocker arm shaft 67 flows through the bore 192 and the lateral bores 200 (see FIG. 8), through the "on-off" gallery 154, the lateral bore 208 and the rocker arm bore 210 to the annular groove 218 in the lash adjuster 76 in the exhaust rocker arm 74a simultaneously through the lateral bore 206 to the lash adjuster 76 in the exhaust rocker arm 74). The pressurized oil flows through the bore 224 in the lash adjuster to the upper chamber 235, thereby developing a hydraulic force which urges the control piston 228 downwardly.

This hydraulic force is opposed by an equal and opposite hydraulic force developed by the pressurized oil in the central chamber 222, which communicates with the constant oil supply gallery 155 through the bores 176 and 178, the annular groove 217, and the lateral bores 219 and 220. Accordingly, the control piston 228 is driven downwardly by the compression spring 233, and the control piston finger 238 displaces the ball check valve 245 away from its valve seat in the inner edge of the surface 254 adjacent the lower entrance to the port 239, thereby overcoming the light biasing force of the spring 243.

The compression spring 250 can drive the socket piston 247 against the snap ring 256 which the cam follower 80 rides upon the secondary exhaust cam base circle 43, and the pressurized oil can flow from the central chamber 222 through the port 239 and fill the expanded chamber 240, but the socket piston 247 is not hydraulically locked in its expanded position because the ball check valve 245 is held off its seat. Thus, when the cam follower 80 rides up the auxiliary opening profile 46, the pushrod 78 drives the socket piston 247 upwardly against the stop 258, thus collapsing the high pressure chamber 240 (and thereby collapsing the lash adjuster 76). The excess oil expelled from the chamber 240 flows through the central chamber 222 and out the lateral bores 219 and 220.

In summary, when the hydraulic lash adjuster is in its "collapsed" state, the socket piston 247 is reciprocated by the compression spring 250 and the pushrod 78, and the pressurized oil flows back and forth through the bores 219, 220, 176 and 178 (see FIG. 9A). Accordingly, the hydraulic lash adjuster 76 has no effect on the timing of the corresponding exhaust valve, which opens and closes at the normal times.

In order to obtain an engine braking effect from the two cylinders associated with the cylinder head 63, when the manual cab switch 100 is closed and the control rack 106 is in its no fuel position to close the limit switch 102, the operator depresses the brake pedal 98 to close the switch 93 and thus energize the solenoid 85. The ball valve 198 is driven against the lower seat 197 to interrupt the communication between the constant oil supply gallery 155 and the "on-off" oil gallery 154, whereupon the pressure of the oil in the upper chamber 226 drops to atmospheric pressure, inasmuch as the upper chamber is communicated to atmosphere through the bores 224, 210, 208, 154, 200, 195 and 204. The upward force developed by the pressurized oil in the center chamber 222 then overcomes the downward biasing force of the spring 233, so that the control piston 228 is driven upwardly against the plug 229. This expells an amount of oil through the lateral bores 204 in the valve body 191 equal to the reduction in volume of the upper chamber 226.

Withdrawal of the control piston finger 238 permits the ball check valve 245 to seat Thus, when the cam follower 80 rides on the secondary cam base circle 43, permitting the spring 250 to drive the socket piston 247 against the snap ring 256, pressurized oil from the central chamber 222 is drawn into the high pressure chamber 240 through the port 239, overcoming the biasing force of the spring 243, but the oil is trapped and locked in the high pressure chamber 240 by the check vlave 245 (see FOG, 9B). Accordingly, the socket piston 247 (and thus the lash adjuster 76) is hydraulically locked in its expanded position, so that the two exhaust valves are opened as the cam followers 80 ride up on the secondary opening profiles 46 near the end of the compression stroke, thus dumping the compressed air so that the energy thereof is dissipated to atmosphere and does not return to the engine during the expansion stroke.

Should the operator desire a greater engine braking effect, he further depresses the brake pedal 98 to close the switches 94, 95 and 96 and obtain the additional braking effect of the pairs of cylinders associated with the cylinder heads 64, 65 and 66, rspectively, as discussed above. Thereafter, additional braking may be obtained by further depressing the brake pedal to activate the conventional service brakes.

FIGS. 11A-D show alternative electrical control circuitry which may be used with the engine braking system of FIGS. 4-10. If a graduated degree of engine braking is not necessary for a particular application, then the electrical control circuitry shown in FIG. 11A may be employed, in which a single switch 260 is mounted so as to be closed when the brake pedal 98 is partially depressed, but before the conventional service brake is actuated. When the manual cab switch 100 is closed and the control rack 106 is in the no-fuel position so as to close the limit swich 102, the closing of the switch 260 simultaneously energizes the solenoids 85-88, thereby to convert all eight cylinders of the V-8 engine to air compressors and thus obtain the maximum engine braking effect.

If all brake pedal actuated switches are omitted, as illustrated in FIG. 11B, then the solenoids 85-88 will be simultaneously energized to obtain the maximum engine braking effect whenever the vehicle overruns the particular setting of the throttle (provided the cab switch 100 is closed), which occurs, for example, when the throttle pedal is released, whereupon the governor drives the control rack 106 to the no-fuel position and closes the limit switch 102. (As discussed above, when idling speed is reached, the governor will automatically move the rack to the idle fuel position, after deactivating the engine braking, thereby preventing stalling of the engine.) With this control system, it will be noted that the lighter degree of "motoring" friction is unavailable, since the engine brake is automatically activated whenever the injection of fuel is stopped (assuming the cab switch is closed).

This system has the advantage of acting as a clutch brake during upshifting of slowing down the countershaft(s) and thus bringing the gear set to be engaged more rapidly into synchronism. On the other hand, difficulty might be encountered in double clutching for downshifting, inasmuch as the counter-shaft(s) will be slowed down by the engine brake before it is deactivated when the throttle is opened to speed up the counter-shaft(s), so that it will take longer to get the countershaft(s) up to synchronized speed. The operator can overcome this difficulty by insuring that the throttle pedal is sufficiently depressed during downshifting, so that the limit switch 102 does not close and the engine brake is not activated.

The control circuitry of FIG. 11C overcomes the down-shifting difficulty of the previous control circuitry by including a series connected manual switch 262 mounted on the shift lever, thereby permitting the operator to deactivate the engine brake during downshifting without having to manipulate the throttle, while retaining the clutch brake effect during up-shifting. Thus, during upshifting the switch 262 is left in its normally closed position, so that the engine brake is automatically activated during upshifting (assuming the cab switch 100 is closed) to slow down the countershaft(s) when the throttle is released to idle position. During downshifting, however, the operator opens the switch 262 to prevent any undesirable engine (clutch) braking during this operation.

In the control system according to FIG. 11D, a normally closed pedal switch 264 is included, thereby eliminating the engine brake during downshifting. When the clutch is disengaged, the switch 264 is opened to deactivate the engine brake, and when the transmission is in neutral and the clutch engaged during double clutching, the throttle will be opened to speed up the countershaft(s) and simultaneously open the limit switch 102 to deactivate the engine brake. On the other hand, during upshifting, when the transmission is in neutral and the clutch engaged during double shifting, the engine brake will be automatically activated to act as a clutch brake and facilitate the shift by slowing down the countershaft(s).

FIG. 12 shows another embodiment of the engine braking system of FIGS. 4-10. In this embodiment, a hydraulic lash adjuster 270, similar in construction and identical in operation with the hydraulic lash adjuster 76 of the previous embodiment, is mounted in the engine block 61 rather than in the exhaust rocker arm. Accordingly, each exhaust valve 72 is reciprocated by a conventional exhaust rocker arm 272, which is identical to the corresponding one of the inlet rocker arms 133 and 133a of the previous embodiment. Each pair of exhaust rocker arms 272 associated with one of the four cylinder heads of the V-8 engine is carried by a rocker arm shaft 274 having a single longitudinal bore 275 therein, the bore 275 being supplied with pressurized oil from the lubrication pump 160. Each rocker arm shaft 274 contains four lateral bores (not shown) which communicate lubricating oil from the longitudinal bore 275 to the two exhaust rocker arms and the two inlet rocker arms carried thereby.

Each exhaust pushrod 78 is received at its upper end in the adjusting screw 148, which is threadedly received in the exhaust rocker arm 272 so as to enable adjustment to provide sufficient lash to accomodate the thermal expansion of the valve train elements, the adjusting screw being locked in the desired position by the locking nut 150. The lower end of the pushrod 78 is received in one end of the lash adjuster 270, the other end of which rides upon the exhaust cam 42.

As before, a graduated degree of engine braking may be obtained by depressing the brake pedal 98 to sequentially close the switches 93, 94, 95 and 96, assuming that the cab switch 100 and the control rack limit switch 102 are closed. The switches 93, 94, 95 and 96 are connected to the solenoids 278, 279, 280 and 281, respectively, which, when energized, activate the valves 283, 284, 285 and 286, respectively, by which the pressurized oil from the pump 160 may be selectively communicated with one portion of the lash adjusters 270, pressurized oil being constantly supplied to the other portion thereof.

Referring now to the enlarged view of one of the pair of lash adjusters 270 associated with the solenoid 278 shown in FIG. 12A, oil under pressure is supplied by the pump 160 through a conduit 290 to a constant oil supply gallery 291, which has branch conduits 291a and 291b communicating with the central chambers 222' of the two-lash adjusters 270 whose operation is controlled by the valve 283. Each lash adjuster 270 reciprocates in a sleeve 293 mounted in the engine block 61. The lash adjuster 270 includes a body 295 having a first annular groove 297 formed in the outer lateral surface thereof, a pair of lateral bores 298 and 299 communicating the groove 297 with the central chamber 222'. The axial length of the groove 297 is such that during the reciprocation of the lash adjuster 270 in the sleeve 293, the chamber 222' is always in communication with the constant oil supply gallery 291 through a bore 300 in the sleeve.

When the solenoid 278 is de-energized, as shown in FIG. 12A, the valve 283 communicates with the oil under pressure in the conduit 290, and a conduit 302 connected thereto, with an "on-off" oil gallery 304, which in turn supplies oil under pressure through the branch conduits 304a and 304b to the chambers 226' of the two hydraulic lash adjusters 270 controlled by the valve 283. In particular, the gallery 304 communicates with each chamber 226' through a bore 306 in the sleeve 293, a groove 307 formed in the outer lateral surface of the body 295, and a lateral bore 308 in the body 295. The axial length of the groove 307 is such that it is always in communication with the "on-off" oil gallery 304.

The compression spring 233' is received in the bore 235' in the control piston 228', and bears against a plug 310 which seals the chamber 226', a seal ring 312 providing a liquid-tight seal between the plug 310 and the body 295. The lower surface of the plug 310 rides on the exhaust cam 42. The socket piston 247' receives in its socket 249' the lower end of the pushrod 78, and the socket piston may reciprocate between the stop 258' and the snap ring 256'.

With pressurized oil in both chambers 222' and 226', the spring 233' drives the control piston 228' upwardly againt the stop 237', so that its finger 238' engages the ball check valve 245' and lifts it from it seat, overcoming the biasing force of the light spring 243'. When the lash adjuster 270 is riding on the secondary exhaust cam base circle 43, the compression spring 250' drives the socket piston 247' against the snap ring 256', but as the lash adjuster rides up the auxiliary opening profile 46, the pushrod 78 collapses the socket piston 247' (and thus the lash adjuster 270), driving the excess oil out of the chamber 240', inasmuch as the ball check valve 245' is held off its seat by the control piston 228'.

When the injection pump switch 102 closes, sensing and verifying a zero fuel condition, and the switch 93 is closed by the brake pedal 98 to energize the solenoid 278 (assuming the cab switch 100 is closed), the valve 283 is actuated to de-couple the "on-off" gallery 304 from the pressurized oil in the conduit 290, and the pressure of the oil in the chamber 226' and in the gallery 304 drops to atmospheric pressure, due to the communication of the gallery 304 with atmosphere through the valve 283 and a vent conduit 314. The pressure of the oil in the chamber 222' then overcomes the force of the spring 233', and so the control piston 228' is driven downwardly against the plug 310, thus disengaging the finger 238' from the ball check valve 245'. The oil driven from the chamber 226' is relieved through the conduits 304 and 314.

Accordingly, when the plug 310 next rides on the secondary exhaust cam base circle 43, the spring 250' drives the socket piston 247' against the snap ring 256', drawing additional pressurized oil into the high pressure chamber 240'. This oil is locked in the chamber 240' by the ball check valve 245', which is urged against its seat by the biasing spring 243', and so the socket piston 247' (as well as the lash adjuster 270) is locked in its expanded position.

The lash being removed from the exhaust valve mechanism (except for the lash 48 shown in FIG. 4A), the exhaust valves 72 associated with the solenoid 278 are opened by the auxiliary opening profile 46 near the end of the compression stroke to provide ening braking. Should a greater engine braking effect be desired, the brake pedal is further depressed to energize the desired ones of the solenoids 284-286, as discussed above.

In the above-described operation of the engine braking system, it will be recalled that each lash adjuster 76 is positioned in its rocker arm 74 (see FIG. 5), and each adjusting screw 148 is positioned in its rocker arm 272 (see FIG. 12), when the valve train elements are cold to accommodate the thermal expansion thereof. Accordingly, the time at which the exhaust valve opens to dump the compressed air for engine braking varies in accordance with the point along the auxiliary opening profile 46 at which the prevailing lash 48 is taien up, and this is determined by the degree of thermal expansion of the valve train elements, upon which the amount of the lash 48 depends.

The time at which each exhaust valve is opened to dump the compressed air may be maintained at a desired point by removing all lash during the engine braking. With zero lash, each exhaust valve opens as soon as the associated cam followr is raised by the initial portion of the auxiliary opening profile 46, regardless of any variation in the exhaust valve train length due to thermal effects.

The zero lash is obtained by positioning each lash adjuster 76, or adjusting screw 148, in its rocker arm 74, or 272, when the valve train elements are cold, so that the corresponding cam follower engages the secondary exhaust cam base circle 43 when the associated hydraulic lash adjuster is expanded. Any extension of the corresponding valve train length due to thermal expansion or other loading effects is accommodated by a corresponding small leakage of oil from the high pressure chamber 240 or 240' past the socket piston 247 or 247'. Such leakage is so slow that movement of the socket piston is negligible during the valve event.

If the valve train length should contract from thermal or other effects, the zero lash condition will be maintained (during engine braking) during the unloading or valve closing portion of the lift cycle when the spring 250 or 250' drives the socket piston 247 or 247' against the secondary base circle 43, drawing pressurized oil from the chamber 222 or 222' into the high pressure chamber 240 or 240', where the oil is trapped and locked by the ball check valve 245 or 245'.

Accordingly, when the engine braking apparatus is operated with all lash removed, the socket piston 247 and 247' (and thus the hydraulic lash adjuster 76 or 270, respectively) is hydraulically locked in an expanded condition in which the associated cam follower engages the secondary base circle 43, even though the socket piston may not engage the snap ring 256 or 256'.

FIGS. 13-21 relate to another embodiment of the variable valve timing system in which the starting of a relatively low compression ratio internal combustion engine is enhanced by increasing the effective compresson ratio during starting. As shown in FIG. 13, the inlet valve opening sector 320 indicats that in a typical compression-ignition engine the inlet valve opens approximately 20.degree. before the top dead center postion of the engine crank during the latter part of the exhaust stroke and before the exhaust valve closes (see the opening sector 321) to promote scavening, and the inlet valve closes approximately 40.degree. after bottom dead center during the initial portion of the compression stroke, with the normal valve tappet clearance or lash to accommodate the thermal expansion of the valve train components. The actual beginning of compression occurs when the inlet valve closes, and so under normal operation, the effective compression ratio is somewhat lower than the numerical or theoretical compression ratio, which could only be obtained if the inlet valve closed at bottom dead center.

The inlet valve timing is shown in FIG. 14. during normal operation with a standard lash indicated at 322, the normal timng corresponds to the inlet valve opening duration shown at 320 in FIG. 13. If excess clearance or lash is provided in the inlet gear train as indicated at 324, the inlet valve opening function is reduced, retarding the opening point to somewhat after top dead center and advancing the closing point to bottom dead center. This reduced inlet valve opening duration caused by excess clearance is illustrated by the opening sector 326 in FIG. 15.

FIG. 16 shows the cylinder pressure characteristics during the compression and expansion strokes when the inlet valve normally closes somewhat after bottom center (see curve 328) and also when the inlet valve closes (and thus the actual beginning of compression starts) at bottom center, as seen in the curve 330. It is apparent that the peak compression is higher when the inlet valve closes at bottom dead center as compared with the peak pressure attained when the inlet valve normally closes after bottom dead center.

FIG. 17 shows the positions of a control connecting rod 336 and piston 337 in a cylinder 338 when the inlet valve closes normally somewhat after bottom dead center during the initial portion of the compression stroke. During such normal operation, the actual compression ratio depends upon the volume swept by the piston during the compression stroke after the inlet valve closes, indicated at 339, and the clearance volume which exists between the piston 337 and the cylinder head when the piston is at top center, such clearance volume being indicated at 340.

FIG. 18 shows the position of the crank, connecting rod and piston when the inlet valve closes at bottom center. Hence it will be noted that the effective compression volume, indicated at 341, is greater than the effective compression volume indicated at 339 under normal conditions, although the clearance volume 340 in the two cases is identical, thereby providing a higher peak compression pressure and a higher actual or effective compression ratio when the closing of the inlet valve is advanced to bottom dead center.

As is well known, reducing the compression ratio of an internal combustion engine reduces the peak cylinder firing pressures and thus permits a higher engine output to be obtained at the same peak firing pressures as are developed in a higher compression ratio engine of lower output. A conventional internal combustion engine having a rated or theoretical compression ratio of 15:1, for example, actually has an effective compression ratio of approxiately 14:1 due to the fact that the inlet valve closes somewhat after bottom center during the initial portion of the compression stroke, as discussed above. If excess clearance or lash is introduced into the inlet valve gear train in order to advance the closing of the inlet valve to bottom center during starting, the effective compression ratio of the "15:1" compression ratio engine will be increased to approximately 15:1, for example, thereby enhancing the starting of the compression-ignition engine.

Alternatively, the performance of the engine could be further increased by redesigning it to provide a still lower effective compression ratio of 13:1, for example, during normal operation when the inlet valve closes somewhat after bottom center, and then by introducing excess clearance or lash in the inlet valve gear train during starting, the actual compression ratio would be approximately 14:1. Accordingly, the startng of the modified low compression ratio engine would equal that of a conventional enging having a rated compression ratio of 15:1, but an actual effective ratio of only 14:1.

The variable valve timing system for selectively introducing excess clearance into the inelt valve gear train in order to increase the effective compression ratio of a relatively low compression ratio engine during starting is shown in FIG. 19, as applied to the same basic V-8 engine used to illustrate the engine braking system described above. Accordingly, identical components of the engine in these two embodiments bear the same reference numerals and need not be described again in detail.

Each inlet valve 345 is driven by a corresponding one of the pair of inlet rocker arms 346 and 346a associated with each of the cylinder heads 63-66. The inlet rocker arms 346 and 346a are identical to the exhaust rocker arms 74 and 74a described above in connection with the engine braking system and illustrated in FIGS. 7-9.

The opposite end of each inlet rocker arm mounts the hydraulic lash adjuster 76, which has been described in detail above and is illustrated in FIGS. 9 and 10. Each lash adjuster 76 receives the upper end of a pushrod 348, the lower end of which is received in a cam follower 350, which engages a conventional inlet cam 352 formed on the camshaft 55, which in turn is conventionally driven by the crankshaft 56.

The hydraulic lash adjusters 76 are threadedly received in the rocker arms 346 and 346a so that the lash adjuster may be positioned in the rocker arm when the inlet valve train components are cold so as to provide a standard lash indicated at 322 in FIG. 14 to accomodate the thermal expansion of the valve gear train elements. Each lash adjuster is locked in the desired position by a lock nut 82.

Each of the cylinder heads 63-66 mounts a rocker arm shaft 355, upon which are mounted for rocking motion the inlet rocker arms 346 an 346a, as well as a pair of exhaust rocker arms 357 and 357a, which are identical to the inlet rocker arms 133 and 133a described above in connection with the engine braking system and illustrated in FIGS. 5, 7 and 8.

Each rocker arm shaft 355 includes a constant oil supply gallery 360, closed at each end by a plug 361, and an "on-off" gallery 363, plugged at each end by a plug 364 (see FIG. 19A). As before, oil under pressure is supplied through the conduit 158 to the constant oil supply gallery 360 from the oil pump 160, which draws oil from the engine sump 161 through the conduit 162, the condut 158 communicating with the passageway 360 through the bore 164 in the rocker arm shaft supporting bracket 131. The bore 164 includes a threaded connection diameter portion 165 which threadedly receives a plug 166, which extends into and is snugly received in a lateral bore 366 formed in the rocker arm shaft 355 and communicating with the passageway 360. The plug 166 locks the rocker arm shaft against rotational or axial displacement with respect to the rocker arm shaft supporting brackets, and a bore 170 extending through the plug communicates the conduit 158 with a constant oil supply gallery 360.

Oil under pressure flows from the oil gallery 360 through a pair of lateral bores 368 and 369 to lubricate the rocking motion of the exhaust rocker arms 357 and 357a, and also flows through a pair of lateral bores 370 and 371 to lubricate the rocking motion of the inlet rocker arms 346 and 346a and to supply oil under relatively constant pressure through a lower bore 373 in the inlet rocker arms to the hydraulic lash adjusters 76.

As in the previous embodiment, the inlet and exhaust rocker arms 346, 346a, 357 and 357a, are grooved and bored (see the bore 375 in FIG. 19) so that sufficient oil from the constnat supply gallery 360 drips over the noses of the rocker arms and thus lubricates the slipper faces thereof.

Each rocker arms shaft 355 is formed with a lateral bore 377 to threadedly receive the conrol valve 190 by means of which oil under pressure may be communicated from the oil gallery 360 to the "on-off" oil gallery 363, or in the oil in the "on-off" gallery may be exposed to atmospheric pressure, as discussed above, The solenoid 85 is mounted on the body of the control valve 190 and includes the armature 193 for displacing the ball valve element 198 of the valve 190. When the solenoid 85 is de-energized, pressurized oil from the gallery 360 is communicated through the "on-off" gallery 363, through a pair of lateral bores 380 and 381 in the rocker arm shaft 355, and through an upper bore 383 in each of the inlet rocker arms to the hydraulic lash adjuster 76 mounted therein.

The lower bore 373 and the upper bore 383 in each inlet rocker arm communicate with the central chamber 222 and the upper chamber 226, respectively, in the lash adjuster 76, (see FIG. 10), and when the solenoid 85 is de-energized, pressurized oil is supplied above and below the control piston 228, which is then driven downwardly by the compression spring 223 to displace the ball check valve 245 from its seat, thereby collapsing the hydraulic lash adjuster 76, as explained above.

When the solenoid 85 is energized, the control valve 190 is actuated to seal the on-off gallery 363 from the constant oil supply gallery 360, and at the same time expose the oil in the gallery 363 and in the upper chamber 226 of the lash adjuster to atmospheric pressure, whereupon the force developed by the pressurized oil in the central chamber 222 overcomes the force of the compression spring 233 and the control piston 228 is driven upwardly, thereby permitting the ball check valve 245 to seat and hydraulically lock the socket piston 247 (and thus the lash adjuster 76) in its expanded position.

When the operator wishes to start the relatively low compression ratio V-8 engine illustrated in FIG. 19, he opens a manual switch 385 located in the cab to de-energize the four solenoids 85 associated with the cylinder heads 63-66, thereby collapsing the hydraulic lash adjusters 76 mounted in all of the inlet rocker arms 346 and 346a. This contracts all of the inlet valve gear trains, providing therein the excess clearance indicated at 324 in FIG. 14, so that the closing time of each inlet valve is advanced to bottom dead center to increase the effective compression ratio to the maximum or theoretical value. This maximizes the peak compression pressures attained during the compression stroke and thereby enhances the starting of the compression-ignition engine.

After the engine has warmed up, the operator closes the cab switch 385, thereby simultaneously energizing the four solenoids 85 so as to expand and hydraulically lock each hydraulic lash adjuster 76 in its expanded condition, in which the socket piston 247 engages the snap ring 246. This expands each inlet valve train to eliminate the excess clearance 324 and provide merely the standard lash indicated at 322 necessary to accommodate the thermal expansion of the inlet valve train elements. Accordingly, the opening time of each inlet valve is retarded to the normal time during the initial portion of the compression stroke, thereby reducing the compression ratio to the nominal value and thus increasing the performance of the engine.

Operating the engine with excess inlet valve clearances for a prolonged period of time or in the upper engine speed range might damage the inlet valve gear components. Accordingly, provision is made to automatically limit the engine speed to a safe range. As discussed above, the control rack 106 of the fuel injection system 108 is displaced to adjust the fuel delivery by one end of the fulcrum lever 115, which is pivoted on the sliding sleeve assembly 114, the position of which is determined by the opposing forces thereon developed by the governor spring 112 and the flyweights 110. The opposite end of the fulcrum lever 115 is positioned by a control lever 388, which in turn is positioned through a linkage 389 by the throttle pedal 390.

A throttle limiting solenoid 392 is mounted in operative relation to the control lever 388, so that a stop member 394 mounted on the solenoid armature normally extends into the path of travel of the control lever 388, the stop member 394 being biased in its extended position by a spring (not shown). The stop member 394 is mounted so that when extended the control lever 388 is engaged at a position at which the governor limits the engine speed to approximately 1,000 RPM for an engine having an operating speed range of 1,200 to 2,100 RPM, for example. Thus, when the manual switch 385 is opened to collapse the hydraulic lash adjusters 76 and provide the excess clearance 324 to increase the effective compression ratio during starting, the extended stop member 394 limits the travel of the contrl lever 388 to prevent the engine speed from exceeding approximately 1,000 RPM, thus preventing any possible damage to the inlet valve gear components.

After the engine has warmed up, the operator closes the manual switch 385 to energize the solenoids 85 and thus expand the lash adjusters 76 (and thereby the inlet valve trains), thereby removing the excess clearance 324, whereupon the inlet valves close during the initial portion of the compression stroke and thereby reduce the effective compression ratio. At the same time, the throttle limiting solenoid 392 is energized to retract the stop member 394 from the path of travel of the control lever 388, overcoming the armature biasing force. Thus, when the standard lash 322 exists in the inlet valve train (providing the lower effective compression ratio), the full throttle operation of the engine is available.

It is to be understood that the engine braking system illustrated in FIGS. 12 and 12A could be included in the variable effective compression ratio engine described above in connection with FIG. 19. Alternatively, the engine braking system illustrated in FIG. 5 could be used in a low compression ratio engine, the effective compression ratio of which is increased for starting by introducing excess clearance into the inlet valve trains with hydraulic lash adjusters mounted in the block and constructed as the lash adjuster 270 in FIG. 12A. Such lash adjusters would engage the inlet cams 352, and the control valve actuating solenoids 278 would be energized by the electrical control circuitry shown in FIG. 19.

FIG. 20 illustrates an arrangement in which the present variable valve timing system is employed in a V-8 low compression ratio compression-ignition engine to provide a modulated engine braking effect available from four of the cylinders and to enhance starting by increasing the effective compression ratio of the other four cylinders, all of the hydraulic lash adjusters being mounted in the rocker arms.

The cylinder heads 63 and 64 mount the identical rocker arm shafts 67 and 68, respectively, described above and illustrated in FIGS. 7 and 8. These rocker arm shafts mount the conventional inlet rocker arms 133 and 133a, and the exhaust rocker arms 74 and 74a described above and illustrated in FIGS. 7 and 9. Each of the exhaust rocker arms mounts the hydraulic lash adjuster 76, which may be expanded to render effective the auxiliary opening profile 46 of the exhaust cam 42 to open the exhaust valve and effect engine braking, when the corresponding one of the identical solenoids 85 and 86 are energized by depressing the brake pedal to close the switches 93 and 94, respectively, assuming that the switches 100 and 102 are closed, as described above in detail. This permits the operator to obtain the degrees of engine braking represented by the curves 123 and 124 in FIG. 6, before activating the service brakes, if necessary.

The cylinder heads 65 and 66 mount the rocker arm shafts 355, described above and illustrated in FIG. 19A. These rocker arm shafts mount the conventional exhaust rocker arms 357 and 357a, and the inlet rocker arms 346 and 346a described above and illustrated in FIGS. 19 and 19A. Each inlet rocker arm mounts the hydraulic lash adjuster 76, all four of which associated with the cylinder heads 65 and 66 may be simultaneously contracted when the corresponding solenoids 85 are de-energized by opening the manual cab switch 385. This shifts the points on the profiles of the inlet cams 352 at which the corresponding inlet valves 345 open and close, because of the excess clearance 324 introduced in the inlet valve train, as described above in detail. In particular, the inlet valve closing points are advanced to bottom dead center, thereby maximizing the effective compression ratio to facilitate starting.

Opening the cab switch 385 also de-energizes the throttle limiting solenoid 392, whereupon the spring-biased stop member 394 (see FIG. 19) extends into the path of travel of the control lever 388, thereby limiting the engine speed while the excess clearance 324 is in the inlet valve train and preventing any possible damage to the inlet valve train components. After the engine has warmed up, the cab switch 385 is closed to establish the normal lash 322, and thus the lower compression ratio in the cylinders associated with the cylinder heads 65 and 66, and to retract the stop member 394 to permit full throttle operation.

FIG. 21 illustrates another arrangement in which the variable valve timing system incorporates a modulated engine braking effect and a variable effective compression ratio for enhanced starting in a low compression ratio compression-ignition engine, with all of the hydraulic lash adjusters being mounted in the rocker arms. In this arrangement, however, the engine braking and variable effective compression ratio effects may be obtained from all eight cylinders of a V-8 engine.

Each of the cylinder heads 63-66 mounts a rocker arm shaft 400, upon which are mounted for rocking motion the pair of exhaust rocker arms 74 and 74a and the pair of inlet rocker arms 346 and 346a, each rocker arm mounting a hydraulic lash adjuster 76. Each rocker arm shaft 400 contains a constant oil supply gallery 402 and a pair of "on-off" galleries 403 and 404. The galleries 402, 403 and 404 are plugged at each end by a plug 405, 406 and 407, respectively.

The "on-off" gallery 403 communicates through a pair of lateral bores 410 in the rocker arm shaft with the corresponding upper bores 210 in the exhaust rocker arms 74 and 74a (see FIG. 9), and thus with the upper chambers 226 of the hydraulic lash adjusters 76 mounted therein. A passageway 412 between the oil galleries 402 and 403 receives a control valve 190, by means of which the pressure of the "on-off" gallery 403 (and thus the lash adjuster chambers 226) is selectively varied between atmospheric pressure and the pressure established in the constant oil supply gallery 402 by the oil pump 160 through the conduit 158, the bore 164 in the bracket 131 and a lateral bore 414 in the rocker arm shaft 400, as described in detail in connection with FIG. 8.

A pair of lateral bores 416 in the shaft 400 communicate the constant supply gallery 402 with the lower bores 178 in the corresponding exhaust rocker arms 74 and 74a (see FIG. 9), and thus with the central chambers 222 of the associated lash adjusters 76.

The control valves 190 in the four rocker arm shaft passageways 412 may be actuated to seal the corresponding galleries 403 from the galleries 402 and expose the corresponding lash adjuster upper chambers 226 to atmospheric pressure and expand these exhaust rocker arm lash adjusters to convert the corresponding cylinders to compressors, as the corresponding solenoids 85-88 are energized by the brake pedal switches 93-96, assuming that the switches 100 and 102 are closed. This will provide the graduated engine braking effect illustrated by the curves 123-126 in FIG. 6.

The "on-off" gallery 404 communicates through a pair of lateral bores 420 in the rocker arm shaft with the corresponding upper bores 383 in the inlet rocker arms 346 and 346a (see FIG. 19), and thus with the upper chambers 226 of the hydraulic lash adjusters 76 mounted therein. A passageway 422 between the galleries 402 and 404 receives a control valve 190, by means of which the pressure of the "on-off" gallery 404 (and thus the corresponding lash adjuster chambers 226) is selectively varied between atmospheric pressure and the pressure of the constant supply gallery 402.

A pair of lateral bores 424 in the shaft 400 communicate the gallery 402 with the lower bores 373 in the corresponding inlet rocker arms 346 and 346a, and thus with the central chambers 222 of the associated lash adjusters 76.

The control valves 190 in the four rocker arm shaft passageways 422 may be simultaneously actuated to collapse the lash adjusters 76 in all the inlet rocker arms 346 and 346a when the cab switch 385 is opened to de-energize the four solenoids 85a mounted on the rocker arm shafts 400, to introduce excess clearance in the inlet valve trains and increase the effective compression ratio to enhance starting, as explained above in detail in connection with FIG. 19. Opening the switch 385 also de-energizes the throttle limiting solenoid 392 to limit the maximum engine speed while there is excess clearance in the inlet valve trains, thus preventing possible damage to the inlet valve gear components.

FIGS. 22-24 relate to another embodiment of the variable valve timing system for improving the performance of an internal combustion engine throughout the operating speed range by optimizing the breathing in the lower and higher portions of the operating speed range.

Each inlet cam 430 includes adjacent the cam base circle 431 a pair of contours 432 and 433 (see FIG. 22B) which comprise auxiliary opening and closing profiles, respectively, which are effective during the high speed portion of the engine operating speed range, as will be explained below. The high speed opening and closing profiles 432 and 433 are between the cam base circle 431 and the opening and closing profiles 434 and 435, respectively, which are effective during the low speed portion of the operating speed range.

Similarly, each exhaust cam 438 includes adjacent the cam base circle 439 a pair of auxiliary opening and closing profiles 440 and 441 (see FIG. 22C), which are effective during the high speed portion of the operating speed range, and which are between the cam base circle 439 and the opening and closing profiles 442 and 443, respectively, which are effective during the low speed portion of the operating speed range.

FIG. 22A shows the cam profiles, as well as the lash and valve lift obtained at low and high speeds. When the inlet and exhaust valve gear trains are expanded during high speed operation so as to provide only the lash indicated at 445 to accomodate changes in valve train length due to temperature variations, the high speed opening and closing profiles are effective to advance the opening of the inlet and exhaust valves and to retard the closing thereof. This increases the valve opening duration and causes the additional valve lift indicated at 447 for the inlet and exhaust valves, thereby providing additional breathing during the high speed portion of the operating speed range.

When additional valve lash is introduced by collapsing the inlet and exhaust valve gear trains so as to provide the total lash indicated at 449, the high speed opening and closing profiles are rendered ineffective, and the valve timing is determined by the low speed opening profiles 434 and 442 and closing profiles 435 and 443. This reduces the valve opening duration and the valve lift to provide the appropriate breathing for the low speed portion of the operating speed range.

The variable valve timing system employing the inlet and exhaust cams 430 and 438, respectively, to optimize breathing and improve the engine performance throughout the useful operating speed range is shown in FIG. 23. In the illustrative embodiment, the valve timing system is employed with the same basic V-8 engine described in the previous embodiments, and so identical parts bear the same reference numerals.

Each of the cylinder heads 63-66 mounts a rocker arm shaft 450 (see also FIG. 23A), which in turn mounts for rocking motion two exhaust rocker arms 74 and 74a and two inlet rocker arms 346 and 346a. A hydraulic lash adjuster 76 is mounted in each of the rocker arms, and a control valve operating solenoid 85 is mounted on each rocker arm shaft 450. The construction and operation of the lash adjuster, the solenoid and these rocker arms have been described in detail in connection with the previous embodiments.

Each rocker arm shaft 450 has a constant oil supply gallery 452, plugged at each end by a plug 453, and an "on-off" gallery 455, plugged at each end by a plug 456. Pressurized oil is supplied to the gallery 452 from the oil pump 160 through the conduit 158, the bore 164 in the bracket 131, and a lateral bore 458 in the rocker arm shaft 450. Four lateral bores 460-463 constantly supply pressurized oil from the constant supply gallery 452 through the lower bores 373 and 178 in the inlet and exhaust rocker arms, respectively, to the central chambers 222 in the hydraulic lash adjusters 76 mounted therein.

A threaded bore 465 in the shaft 450 receives the control valve 190, which may be actuated by the solenoid 85 to control the communication between the galleries 452 and 455 and to selectively expose the "on-off" gallery 455 to atmospheric pressure, as explained above in connection with the previous embodiments. Four lateral bores 467-470 in the shaft 450 communicate the "on-off" gallery 455 through the upper bores 383 and 210 in the inlet and exhaust rocker arms, respectively, with the upper chambers 226 in the lash adjusters 76 mounted therein.

The four solenoids 85 are simultaneously energized from the vehicle battery 104 when an engine speed responsive switch 472 is closed. The switch 472 is actuated by a conventional speed responsive device 474 including a switch engaging plunger 475 having a collar 476. A spring 477 bears against one side of the collar and biases the plunger 475 in the direction for opening the switch 472. A plurality of flyweights 478 are driven by the engine through conventional means (not shown), and as the weights revolve, centrifugal force tends to move them outwardly about their pivots 480. The flyweight noses 482 bear against the side of the collar 476 remote from the spring 477, and outward movement of the flyweights 478 drives the plunger 475 in the direction to close the switch 472.

During the low speed portion of the engine operating speed range the centrifugal force on the flyweights 478 is not sufficient to close the switch 472, the solenoids 85 are de-energized, and the hydraulic lash adjusters 76 are collapsed. Accordingly, the relatively large amount of lash indicated at 449 (see FIG. 22A) exists in the inlet and exhaust valve trains, and the shorter valve opening duration and smaller valve lift obtain which are appropriate for low speed operation.

When the engine reaches a predetermined speed at which the valve lift and opening duration are to be increased to provide additional breathing in the high speed range, the force of the flyweights 478 overcomes that of the spring 477, and the switch 472 is closed to energize the four solenoids 85. The lash adjusters 76 are expanded and locked, thereby extending the inlet and exhaust valve gear trains and reducing the lash to that indicated at 445. The high speed opening and closing profiles 432, 440, 433 and 441 are rendered effective, thereby increasing the opening duration and lift of the inlet and exhaust valves, and thus increasing the breathing, volumetric efficiency and performance of the engine in the upper operating speed range.

FIG. 24 illustrates the manner in which the performance of the engine is improved, by showing the volumetric efficiency characteristic. Thus, the curve 485 shows the volumetric efficiency characteristic obtaining when the hydraulic lash adjusters 76 are collapsed and the relatively large amount of lash indicated at 449 exists in the inlet and exhaust valve trains. Under these conditions the valve opening and closing times are determined by the low speed opening and closing profiles 434, 442, 435 and 443. This valve timing maximizes the volumetric efficiency during the low speed portion of the operating speed range.

When the predetermined engine speed indicated at 486 is reached, the speed responsive switch 472 is closed, energizing the solenoids 85 to expand the hydraulic lash adjusters, and thus the inlet and exhaust valve trains, thus reducing the lash to that shown at 445. This renders the high speed opening and closing profiles 432, 440, 433, and 441 effective, thereby increasing the valve opening duration and valve lift to provide the volumetric efficiency characteristic shown by the curve 487, in which the volumetric efficiency is maximized during the high speed portion of the operating speed range.

The speed at which the valve timing is shifted depends upon the characteristics of the particular engine and locations at which the low speed and high speed opening and closing profiles are formed in the cams. For an engine having a useful operating speed range of 1,200-2,400 RPM, for example, the valve timing shift speed might be approximately 1,900 RPM.

FIG. 25 illustrates an arrangement in which the variable valve timing system incorporates a modulated engine braking effect with a system for varying the valve timing in accordance with engine speed in order to improve the engine performance throughout the operating speed range. This system is incorporated in the same basic V-8 engine used to illustrate the previous embodiments.

Each of the cylinder heads 63-66 mounts the rocker arm shaft 400, upon which are mounted the exhaust rocker arms 74 and 74a and the inlet rocker arms 346 and 346a, as in the embodiment described above in connection with FIG. 21. As before, the exhaust cams 42 drive the exhaust rocker arms, the hydraulic lash adjusters of which are controlled by the "on-off" gallery 403 from the corresponding one of the brake pedal switches 93-96 to provide the graduated engine braking effect illustrated by the curves 123-126 of FIG. 6.

The lash adjusters 76 mounted in the inlet rocker arms 346 and 346a are simultaneously controlled by the "on-off" gallery 404 from the four solenoids 85a, as in the FIG. 21 embodiment, but the inlet rocker arms are driven by the inlet cams 430 and the solenoids 85a are energized by the speed responsive switch 472, as in the FIG. 23 embodiment. Accordingly, the timing of the inlet valves is automatically adjusted in accordance with the engine speed to improve the engine performance throughout the useful operating speed range.

It is to be understood that the above-described embodiments are susceptible to modifications, substitutions and changes by those skilled in the art without departing from the spirit and scope of the invention. For example, pneumatic or hydraulic control circuitry could be substituted for part or all of the electrical control circuitry, and fluidic control devices could be substituted for the illustrated switching and valving. Also, the contol valving for the hydraulic lash adjusters could be mounted in other rocker arm support structures, such as the rocker arm shaft supporting brackets, or the individual ball pivots supporting the rocker arms where there is no rocker arm shaft, for example.

Similarly, the lash adjuster control valving could be mounted independently of, but adjacent to, the rocker arm support structure, external conduits coupling the control valving with the lash adjusters through the rocker arm shaft, or directly through the rocker arms, a flexible hose or a slip joint being provided between each rocker arm and the external conduit. Also, such external control valving and conduits could be employed to control the lash adjusters in the inlet rocker arms to vary the timing of the inlet valves to enhance starting or to increase engine performance, while the lash adjusters in the exhaust rocker arms are controlled by control valving and conduits in the rocker arm shaft, for example.

Furthermore, equivalent valve elements could be substituted for the ball valve element in the control valves and the ball check valve in the hydraulic lash adjusters. Also, other conventional adjustable, normally open centrifugal switches could be substituted for the speed responsive switch 472, such as an off-center weighted spring leaf carrying the necessary contact points or activating a microswitch, for example. In addition, the face of each piston could be suitably recessed to accommodate the additional valve lift when the valve gear trains are expanded, thereby preventing engagement between the valve head and the piston, rather than recessing the valve seat inserts 73.

It is also to be understood that the hydraulic lash adjusters could be independent of the gear train by which the valves are reciprocated during normal operation, the lash adjusters being mounted in operative relation to auxiliary opening and closing profiles which are formed on auxiliary cams remote from the cams which determine the normal valve timing. In such modification, the lash adjusters would be expanded to render the auxiliary opening and closing profiles effective for engine braking, to enhance the starting of a relatively low compression ratio engine, or to adjust the valve timing in accordance with engine speed to improve engine performance. Accordingly, all such substitutions and modifications are to be included in the scope of the invention as defined by the following claims.

Claims

We claim:

1. A variable valve timing system for an internal combustion engine having a crankshaft and at least one cylinder, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, and at least one valve for each cylinder, comprising at least one cam means driven by the crankshaft, the cam means including a first profile for actuating a corresponding valve in a first timed relation to the rotation of the crankshaft and a second profile for actuating at least one of the valves in a second timed relation to the rotation of the crankshaft, the cam means for actuating any valve in two timed relations to the rotation of the crankshaft including a cam having the first and second profiles formed thereon circumferentially spaced from each other, a valve gear train coupling each valve to a corresponding one of the cam means, means for expanding at least one of the valve gear trains from a collapsed condition in which the first profile is effective for actuating the corresponding valve in the first timed relation to the rotation of the crankshaft to an expanded condition in which the second profile is effective for actuating the corresponding valve in the second timed relation to the rotation of the crankshaft, each valve capable of being actuated in the two timed relations to the rotation of the crankshaft being coupled to the first and second profiles of the corresponding cam by a single valve gear train, and means for locking each expanded valve gear train in the expanded condition, whereby each valve actuated by an expandable valve gear train is actuated in one of two predetermined timed relations to the rotations of the crankshaft corresponding to the collapsed and the locked expanded conditions of the corresponding valve gear train, each expanding means including means forming a hydraulic chamber having an inlet port, a piston reciprocally disposed in the chamber, and means for supplying pressurized hydraulic fluid through the inlet port into the chamber, and the locking means including check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber, and means for biasing the check valve means to seal the chamber to prevent hydraulic fluid in the chamber from exiting therefrom.

2. A variable valve timing system for an internal combustion engine having a crankshaft and at least one cylinder, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, and at least one valve for each cylinder, comprising at least one cam means driven by the crankshaft, the cam means including a first profile for actuating a corresponding valve in a first timed relation to the rotation of the crankshaft and a second profile for actuating at least one of the valves in a second timed relation to the rotation of the crankshaft, the cam means for actuating any valve in two timed relations to the rotation of the crankshaft including a cam having the first and second profiles formed thereon circumferentially spaced from each other, a valve gear train coupling each valve to a corresponding one of the cam means, means for expanding at least one of the valve gear from a collapsed condition in which the first profile is effective for actuating the corresponding valve in the first timed relation to the rotation of the crankshaft to an expanded condition in which the second profile is effective for actuating the corresponding valve in the second timed relation to the rotation of the crankshaft, each valve capable of being actuated in the two timed relations to the rotation of the crankshaft being coupled to the first and second profiles of the corresponding cam by a single valve gear train, each expanding means including means forming a hydraulic chamber having an inlet port, a piston reciprocally disposed in the chamber, and means for supplying pressurized hydraulic fluid through the inlet port into the chamber, means for locking each expanded valve gear train in the expanded condition, whereby each valve actuated by an expandable valve gear train is actuated in one of two predetermined timed relations to the rotation of the crankshaft corresponding to the collapsed and the locked expanded conditions of the corresponding valve gear train, the locking means including check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber through the inlet port, and means for biasing the check valve means to seal the inlet port to prevent the hydraulic fluid from exiting therethrough, and also including means for biasing the piston in the direction to expand the associated valve gear train, and control means for selectively disabling the check valve means and preventing the hydraulic fluid from being trapped in the chamber, whereby when the check valve means is disabled, the piston is able to reciprocate in the chamber, the corresponding second profile is rendered ineffective, and the corresponding valve is actuated in the first timed relation to the rotation of the crankshaft.

3. A variable valve timing system for an internal combustion engine having a crankshaft and at least one cylinder, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, and at least one valve for each cylinder, comprising at least one cam means driven by the crankshaft, the cam means including a first profile for actuating a corresponding valve in a first timed relation to the rotation of the crankshaft and a second profile for actuating at least one of the valves in a second timed relation to the rotation of the crankshaft, a valve gear train coupling each valve to a corresponding one of the cam means, means for expanding at least one of the valve gear trains to render the second profile effective for actuating the corresponding valve in the second timed relation to the rotation of the crankshaft, each expanding means including means forming a hydraulic chamber having an inlet port, a piston reciprocally disposed in the chamber, and means for supplying pressurized hydraulic fluid through the inlet port into the chamber, means for locking each expanded valve gear train in an expanded condition, the locking means including check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber through the inlet port, means for biasing the piston in the direction to expand the associated valve gear train, means for biasing the check valve means to seal the inlet port to prevent the hydraulic fluid from exiting therethrough, and control means for selectively disabling the check valve means and preventing the hydraulic fluid from being trapped in the chamber, whereby when the check valve means is disabled, the piston is able to reciprocate in the chamber, the corresponding second profile is rendered ineffective, and the corresponding valve is actuated in the first timed relation to the rotation of the crankshaft, the control means including means forming a control chamber, a control piston reciprocally disposed in the control chamber, the control piston including a finger adapted to displace the check valve means into spaced non-sealing relation to the inlet port to permit hydraulic fluid to exit therethrough, means for biasing the control piston to displace the check valve means, and means for supplying pressurized hydraulic fluid to the portion of the control chamber intermediate the control piston and the check valve means to develop a force which urges the control piston away from the check valve means.

4. The system according to claim 3, wherein the cam means for actuating any valve in two timed relations to the rotation of the crankshaft includes a cam having the first and second profiles formed thereon circumferentially spaced from each other.

5. The system according to claim 3, wherein the control means also includes means for supplying hydraulic fluid to the portion of the control chamber on the side of the control piston remote from the check valve means, and means for selectively shifting the pressure of the hydraulic fluid in the remote control chamber portion between substantially atmospheric pressure and substantially the same pressure as that of the pressurized hydraulic fluid in the control chamber portion between the control piston and check valve means.

6. The system according to claim 1, wherein each hydraulic chamber forming means and each corresponding piston comprise a portion of the corresponding valve gear train coupling the corresponding valve with the second cam profiles.

7. The system according to claim 6, wherein each hydraulic chamber forming means and each corresponding piston are mounted for reciprocation in the block of the internal combustion engine.

8. The system according to claim 6, wherein the valve gear trains include rocker arms coupling the valves to at least some of the cam means, and each hydraulic chamber forming means and each corresponding piston are mounted in the rocker arm which actuates the corresponding valve in the first and second timed relations to the rotation of the crankshaft.

9. The system according to claim 8, including at least one rocker arm shaft for mounting the rocker arms for rocking motion, each rocker arm shaft mounting a rocker arm coupled to a valve whose timing may be varied being formed with two passageways, the first passageway being adapted to be communicated with a source of pressurized hydraulic fluid for supplying pressurized hydraulic fluid for lubricating each rocker arm carried by the rocker arm shaft, and the second passageway being adapted to be communicated with the expanding means, control valve means mounted in each rocker arm shaft formed with the two passageways for selectively communicating the second passageway with the first passageway, and for selectively sealing the second passageway from the first passageway and exposing the hydraulic fluid in the second passageway to atmospheric pressure, and means for actuating the control valve means to expand the valve gear trains of the valves whose timing is to be varied.

10. A variable valve timing system for an internal combustion engine having a crankshaft and at least one cylinder, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, and at least one valve for each cylinder, comprising at least one cam means driven by the crankshaft, the cam means including a first profile for actuating a corresponding valve in a first timed relation to the rotation of the crankshaft and a second profile for actuating at least one of the valves in a second timed relation to the rotation of the crankshaft, a valve gear train coupling each valve to a corresponding one of the cam means, each valve gear train including a rocker arm coupling the valve to a corresponding cam means, means for expanding at least one of the valve gear trains to render the second profile effective for actuating the corresponding valve in the second timed relation to the rotation of the crankshaft, each expanding means including means forming a hydraulic chamber having an inlet port, a piston reciprocally disposed in the chamber, and means for supplying pressurized hydraulic fluid through the inlet port into the chamber, means for locking each expanded valve gear train in an expanded condition, the locking means including check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber through the inlet port, each hydraulic chamber forming means and each corresponding piston comprising a portion of the corresponding valve gear train coupling the corresponding valve with the second cam profile and being mounted in the rocker arm which actuates the corresponding valve in the first and second timed relations to the rotation of the crankshaft, at least one rocker arm shaft for mounting the rocker arm for rocking motion, each rocker arm shaft mounting a rocker arm coupled to a valve whose timing may be varied being formed with two passageways, the first passageway being adapted to be communicated with a source of pressurized hydraulic fluid for supplying pressurized hydraulic fluid for lubricating each rocker arm carried by the rocker arm shaft, and the second passageway being adapted to be communicated with the expanding means, control valve means mounted in each rocker arm shaft formed with the two passageways for selectively communicating the second passageway with the first passageway, and for selectively sealing the second passageway from the first passageway and exposing the hydraulic fluid in the second passageway to atmospheric pressure, means for actuating the control valve means to expand the valve gear train of each valve whose timing is to be varied, means forming a control chamber mounted in each rocker arm coupled to a valve whose timing is to be varied, a control piston reciprocally disposed in the control chamber, the control piston including a finger adapted to displace the check valve means into spaced non-sealing relation to the inlet port to permit hydraulic fluid to exit therethrough, means for biasing the control piston to displace the check valve means, a first conduit communicating the portion of the control chamber intermediate the control piston and the check valve means with the first passageway in the rocker arm shaft, and a second conduit communicating the portion of the control chamber on the side of the control piston remote from the check valve means with the second passageway in the rocker arm shaft.

11. The system according to claim 10, including means for biasing the piston in the hydraulic chamber to expand the associated valve gear train, and means for biasing the check valve means to seal the inlet port of the hydraulic chamber.

12. A variable valve timing system for an internal combustion engine having a crankshaft and at least one cylinder, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, and at least one inlet valve and at least one exhaust valve for each cylinder, comprising a plurality of cam means driven by the crankshaft, the cam means including first profiles for actuating the valves in a first timed relation to the rotation of the crankshaft and second profiles for actuating at least one of the valves in a second timed relation to the rotation of the crankshaft, a plurality of valve gear trains coupling the inlet and exhaust valves to corresponding ones of the cam means, means for expanding at least one of the valve gear trains from a collapsed condition in which the first profiles are effective for actuating the corresponding valves in the first timed relation to the rotation of the crankshaft to an expanded condition in which the second profiles are effective for actuating the corresponding valves in the second timed relation to the rotation of the crankshaft, and means for locking the expanded valve gear trains in the expanded condition, whereby the valves actuated by the expandable valve gear trains are actuated in one of two predetermined timed relations to the rotation of the crankshaft corresponding to the collapsed and the locked expanded conditions of the corresponding valve gear trains, each expanding means including means forming a hydraulic chamber having an inlet port, a piston reciprocally disposed in the chamber, and means for supplying pressurized hydraulic fluid through the inlet port into the chamber, and wherein the locking means includes check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber through the inlet port, each hydraulic chamber forming means and each corresponding piston comprising a portion of the corresponding valve gear train coupling the corresponding valve with the second cam profiles, the valve gear trains including rocker arms coupling the valves to at least some of the cam means, and each hydraulic chamber forming means and each corresponding piston being mounted in the rocker arm which actuates the corresponding valve in the first and second timed relations to the rotation of the crankshaft, including at least one rocker arm shaft for mounting the rocker arms for rocking motion, each rocker arm shaft mounting a rocker arm coupled to a valve whose timing may be varied being formed with two passageways, the first passageway being adapted to be communicated with a source of pressurized hydraulic fluid for supplying pressurized hydraulic fluid for lubricating each rocker arm carried by the rocker arm shaft, and the second passageway being adapted to be communicated with the expanding means, control valve means mounted in each rocker arm shaft formed with the two passageways for selectively communicating the second passageway with the first passageway, and for selectively sealing the second passageway from the first passageway and exposing the hydraulic fluid in the second passageway to atmospheric pressure, and means for actuating the control valve means to expand the valve gear trains of the valves whose timing is to be varied, each rocker arm shaft mounting a plurality of rocker arms coupled to valves whose timing may be varied being formed with a third passageway adapted to be communicated with some of the expanding means, the second passageway being adapted to be communicated with the other expanding means, and including additional control valve means for selectively communicating the third passageway with the first passageway, and for selectively sealing the third passageway from the first passageway and exposing the hydraulic fluid in the third passageway to atmospheric pressure, and means for actuating the additional control valve means to expand the corresponding valve gear trains independently of the expansion of the valve gear trains associated with the second passageway, whereby the timing of some of the valves may be shifted to one timed relation to the rotation of the crankshaft upon expansion of the associated valve gear trains, and the timing of other valves may be shifted to a different timed relation upon expansion of the associated valve gear trains.

13. The system according to claim 1, wherein each first cam profile is disposed with respect to the position of the crankshaft such that the corresponding valve is opened and closed at optimum times for enhancing engine performance in the lower engine operating speed range, and each second cam profile is disposed with respect to the position of the crankshaft such that the corresponding valve is opened and closed at optimum times for enhancing engine performance in the upper engine operating speed range, and including engine speed responsive means for controlling the means for expanding the valve gear train to render each second cam profile effective.

14. The system according to claim 1, wherein there is at least one inlet valve for each cylinder and at least one inlet valve may be actuated in response to the first and second cam profiles, each second cam profile being disposed with respect to the position of the crankshaft such that the corresponding inlet valve is closed in response thereto during the initial portion of the compression stroke, and each first cam profile being disposed with respect to the position of the crankshaft such that the corresponding inlet valve is closed in response thereto when the corresponding piston is closer to the bottom dead center position, so that the effective compression ratio is higher when the timing of the corresponding inlet valve is determined by the first cam profile than when the timing is determined by the second cam profile, and including means for actuating the means for expanding the valve gear train to render each second cam profile effective and for collapsing the valve grear train to render the first cam profile effective.

15. The system according to claim 14, including means for limiting the maximum engine speed when the valve gear train is collapsed to render the first cam profile effective, the engine speed limiting means being coupled to the means for actuating the valve gear train expanding means so that the maximum engine speed limiting means is automatically activated when the valve gear train is collapsed.

16. The system according to claim 15, wherein the internal combustion engine is supplied with fuel by a fuel injection system including an engine speed governor and a control element which may be positioned for a desired governed engine speed, and wherein the maximum engine speed limiting means includes a displaceable stop member for limiting the travel of the control element.

17. The system according to claim 1, wherein there is at least one exhaust valve for each cylinder and at least one exhaust valve may be actuated in response to the first and second cam profiles, each second cam profile being disposed with respect to the position of the crankshaft such that the corresponding exhaust valve is opened in response thereto at or near the end of the compression stroke to dump the compressed air in the corresponding cylinder so as to increase the retarding effect of the engine, and including means for actuating the means for expanding the valve gear train to render each second cam profile effective and for collapsing the valve gear train to render the first cam profile effective.

18. A variable valve timing system for an internal combustion engine having a crankshaft and at least one cylinder, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, and at least one valve for each cylinder, comprising at least one cam means driven by the crankshaft, the cam means including a first profile for actuating a corresponding valve in a first relation to the rotation of the crankshaft and a second profile for actuating at least one of the valves in a second timed relation to the rotation of the crankshaft, the cam means for actuating any valve in two timed relations to the rotation of the crankshaft including a cam having the first and second profiles formed thereon circumferentially spaced from each other, a valve gear train coupling each valve to a corresponding one of the cam means, means for expanding at least one of the valve gear trains from a collapsed condition in which the first profile is effective for actuating the corresponding valve in the first timed relation to the rotation of the crankshaft to an expanded condition in which the second profile is effective for actuating the corresponding valve in the second timed relation to the rotation of the crankshaft, each valve capable of being actuated in the two timed relations to the rotation of the crankshaft being coupled to the first and second profiles of the corresponding cam by a single valve gear train, and means for locking each expanded valve gear train in the expanded condition, whereby each valve actuated by an expandable valve gear train is actuated in one of two predetermined timed relations to the rotation of the crankshaft corresponding to the collapsed and the locked expanded conditions of the corresponding valve gear train, at least one exhaust valve capable of being actuated in response to the first and second cam profiles, each second cam profile being disposed with respect to the position of the crankshaft such that the corresponding exhaust valve is opened in response thereto at or near the end of the compression stroke to dump the compressed air in the corresponding cylinder so as to increase the retarding effective of the engine, and including means for actuating the means for expanding the valve gear train to render each second cam profile effective and for collapsing the valve gear train to render the first cam profile effective, the internal combustion engine including fuel supply means for delivering fuel to the cylinders, and including means for sensing the fuel delivery to the cylinders, and means responsive to the fuel delivery sensing means for disabling the means for actuating the valve gear train expanding means while fuel is delivered to the cylinders.

19. The system according to claim 18, wherein the fuel supply means includes a movable member for controlling the fuel delivery to the cylinders, and the means for actuating the valve gear train expanding means is responsive to motive force, and including a source of motive force, means for communicating the actuating means with the source of motive force, and wherein the fuel delivery sensing means and the disabling means include switch means in the communicating means operable to selectively effect and terminate communication between the source of motive force and the actuating means, and means coupling the switch means with the movable member.

20. The system according to claim 19, wherein the fuel supply means includes a fuel injection pump having a control rack, and wherein the switch means is mounted in operative relation to the control rack, whereby the switch means is actuated by the control rack to collapse the valve gear trains to render the first cam profiles effective before fuel is delivered to the cylinders.

21. The system according to claim 20, wherein the fuel supply means includes an engine speed responsive governor coupled to the control rack, whereby the governor prevents stalling of the engine by positioning the control rack to actuate the switch means and disable the actuating means and to deliver fuel to the cylinders.

22. The system according to claim 19, wherein the internal combustion engine is mounted in a vehicle having braking means, and a brake pedal for activating the brake means, and including switch means mounted in operative relation to the brake pedal and connected in the communicating means in series with the source of motive force, the disabling switch means and the actuating means, whereby the valve gear trains may be expanded when the brake pedal is depressed to activate the brake pedal switch means and there is no fuel delivery to the cylinders.

23. A variable valve timing system for an internal combustion engine having a crankshaft and at least one cylinder, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, at least one exhaust valve for each cylinder, and means for supplying fuel to each cylinder, comprising means forming a first cam profile for actuating each exhaust valve in a first timed relation to the rotation of the crankshaft for normal operation of the engine, means forming a second cam profile for opening at least one exhaust valve at or near the end of the compression stroke to dump the compressed air from each corresponding cylinder so as to increase the retarding effect of the engine, means for selectively coupling the second cam profile means with each corresponding exhaust valve, means for sensing the fuel delivery to the cylinders, and means responsive to the fuel delivery sensing means for disabling the coupling means and thereby rendering the second cam profile means ineffective while fuel is delivered to the cylinders.

24. Apparatus for varying the timing of at least one of the valves of an internal combustion engine having a crankshaft and at least one cylinder associated with the valves, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, first cam means for actuating the valves in a first timed relation to the rotation of the crankshaft, second cam means for actuating at least some of the valves in a second timed relation to the rotation of the crankshaft, and a plurality of valve gear trains coupling the valves to the corresponding first and second cam means, comprising means forming a hydraulic chamber having an inlet port and adapted to receive pressurized fluid through the inlet port, a piston reciprocally disposed in the chamber, the hydraulic chamber forming means and the piston being a portion of each valve gear train coupling a valve with the corresponding second cam means, check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber, and means for biasing the check valve means to seal the chamber to prevent hydraulic fluid in the chamber from exiting therefrom.

25. Apparatus for varying the timing of at least one of the valves of an internal combustion engine having a crankshaft and at least one cylinder associated with the valves, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, first cam means for actuating the valves in a first timed relation to the rotation of the crankshaft, second cam means for actuating at least some of the valves in a second timed relation to the rotation of the crankshaft, and a plurality of valve gear trains coupling the valves to the corresponding first and second cam means, comprising means forming a hydraulic chamber having an inlet port and adapted to receive pressurized fluid through the inlet port, a piston reciprocally disposed in the chamber, the hydraulic chamber forming means and the piston being a portion of each valve gear train coupling a valve with the corresponding second cam means, check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber through the inlet port, means for biasing the piston in the direction to expand the volume of the hydraulic chamber, means for biasing the check valve means to seal the inlet port to prevent hydraulic fluid in the chamber from exiting therethrough, and control means for selectively disabling the check valve means and preventing hydraulic fluid in the chamber from being trapped therein, whereby when the check valve means is disabled, the piston is able to reciprocate in the chamber, the corresponding second cam means are rendered ineffective, and the corresponding valve is actuated in the first timed relation to the rotation of the crankshaft.

26. Apparatus for varying the timing of at least one of the valves of an internal combustion engine having a crankshaft and at least one cylinder associated with the valves, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, first cam means for actuating the valves in a first timed relation to the rotation of the crankshaft, second cam means for actuating at least one of the valves in a second timed relation to the rotation of the crankshaft, and a plurality of valve gear trains coupling the valves to the corresponding first and second cam means, comprising means forming a hydraulic chamber having an inlet port and adapted to receive pressurized fluid through the inlet port, a piston reciprocally disposed in the chamber, the hydraulic chamber forming means and the piston being a portion of each valve gear train coupling a valve with the corresponding second cam means, check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber through the inlet port, means for biasing the piston in the direction to expand the volume of the hydraulic chamber, means for biasing the check valve means to seal the inlet port to prevent hydraulic fluid in the chamber from exiting therethrough, and control means for selectively disabling the check valve means and preventing hydraulic fluid in the chamber from being trapped therein, whereby when the check valve means is disabled, the piston is able to reciprocate in the chamber, the corresponding second cam means are rendered ineffective, and the corresponding valve is actuated in the first timed relation to the rotation of the crankshaft, the control means including means forming a control chamber, a control piston reciprocally disposed in the control chamber, the control piston including a finger adapted to displace the check valve means into spaced non-sealing relation to the inlet port to permit hydraulic fluid to exit therethrough, and means for biasing the control piston to displace the check valve means.

27. Apparatus according to claim 26, including means forming a first port communicating with the portion of the control chamber intermediate the control piston and the check valve means adapted to admit pressurized hydraulic fluid thereto, and means forming a second port communicating with the portion of the control chamber on the side of the control piston remote from the check valve means adapted to admit pressurized hydraulic fluid thereto.

28. The system according to claim 3, wherein the hydraulic chamber forming means and the corresponding piston of each expanding means are mounted for reciprocation in the block of the internal combustion engine.

29. The system according to claim 3, wherein the valve gear trains include rocker arms coupling the valves to at least some of the cam means, and the hydraulic chamber forming means and the corresponding piston of each expanding means are mounted externally of the rocker arm which actuates the corresponding valve in the first and second timed relations to the rotation of the crankshaft.

30. The system according to claim 3, wherein the valve gear train includes rocker arms coupling the valves to at least some of the cam means, and the hydraulic chamber forming means and the corresponding piston of each expanding means are mounted in the rocker arm which actuates the corresponding valve in the first and second timed relations to the rotation of the crankshaft.

31. The system according to claim 10, wherein the cam means for actuating any valve in two timed relations to the rotation of the crankshaft includes a cam having the first and second profiles formed thereon circumferentially spaced from each other.

32. The apparatus according to claim 26, wherein the hydraulic chamber forming means and the corresponding piston of each valve gear train coupling a valve with the corresponding second cam means are mounted for reciprocation in the block of the internal combustion engine.

33. The apparatus according to claim 26, wherein the valve gear trains include rocker arms coupling the valves to at least some of the cam means, and the hydraulic chamber forming means and the corresponding piston of each valve gear train coupling a valve with the corresponding second cam means are mounted externally of the rocker arm which actuates the corresponding valve in the first and second timed relations to the rotation of the crankshaft.

34. The apparatus according to claim 26, wherein the valve gear trains include rocker arms coupling the valves to at least some of the cam means, and the hydraulic chamber forming means and the corresponding piston of each valve gear train coupling a valve with the corresponding second cam means are mounted in the rocker arm which actuates the corresponding valve in the first and second timed relations to the rotation of the crankshaft.

35. The apparatus according to claim 26, wherein the first and second cam means for each valve capable of being actuated in two timed relations to the rotation of the crankshaft include a first profile and a second profile formed on a single cam, the first and second profiles being circumferentially spaced from each other.

36. Apparatus for varying the timing of at least one of the valves of an internal combustion engine having a crankshaft and at least one cylinder associated with the valves, a piston reciprocable in each cylinder and connected to the crankshaft for rotation thereof, first cam means for actuating the valves in a first timed relation to the rotation of the crankshaft, second cam means for actuating at least one of the valves in a second timed relation to the rotation of the crankshaft, and a plurality of valve gear trains coupling the valves to the corresponding first and second cam means, comprising means forming a hydraulic chamber having an inlet port and adapted to receive pressurized fluid through the inlet port, a piston reciprocally disposed in the chamber, the hydraulic chamber forming means and the piston being a portion of each valve gear train coupling a valve with the corresponding second cam means, check valve means for trapping the pressurized hydraulic fluid in the chamber and preventing the trapped fluid from exiting the chamber through the inlet port, means for biasing the check valve means to seal the inlet port to prevent hydraulic fluid in the chamber from exiting therethrough, and control means for selectively disabling the check valve means and preventing hydraulic fluid in the chamber from being trapped therein, whereby when the check valve means is disabled, the piston is able to reciprocate in the chamber, the corresponding second cam means are rendered ineffective, and the corresponding valve is actuated in the first timed relation to the rotation of the crankshaft.