Hydraulic valve actuator system

A hydraulic valve actuator system for an internal combustion engine in which an engine driven pump means supplies fluid, at high pressure, sequentially to an N number of outlets, each outlet being connected by a check valve and a fluid delivery conduit to an associated actuator means for a poppet valve in a set of N number of poppet valves for the engine, each actuator means including an actuator plunger reciprocably journaled in an actuator cylinder with one end of the actuator plunger positioned for abutment against the stem of an associated poppet valve that is normally held closed by an associated valve return spring. Each fluid delivery conduit is also connected via a drain conduit to a fluid reservoir used to supply fluid at a predetermined supply pressure to the inlet of the pump means, flow from each delivery conduit through its associated drain conduit being controlled by an associated normally open, solenoid valve connectable to an electric control circuit whereby each solenoid valve can be sequentially energized as a function of engine operating perimeters.

FIELD OF THE INVENTION 
This invention relates to internal combustion engines or similar mechanisms 
and, in particular, to a hydraulic valve actuator system for controlling 
the valve operation in such an engine or similar mechanism. 
DESCRIPTION OF THE PRIOR ART 
In most conventional present day internal combustion engines, breathing of 
the engine, that is both inlet to and exhaust from an engine, is 
controlled by means of poppet valves suitably actuated by means of a valve 
drive train actuated by a camshaft of the engine. 
Various attempts have been made to provide for improved control of valve 
actuation in internal combustion engines by the use of hydraulic 
mechanisms to effect valve operation in lieu of a conventional valve drive 
train. To this end, various mechanisms of the type using a hydraulic 
cylinder and piston combination for effecting valve actuation in an engine 
have been proposed in the prior art, such mechanisms of this type 
appropriately being referred to as hydraulic valve actuator mechanisms or 
systems. 
Differing arrangements of one form of such hydraulic valve actuator systems 
are disclosed, for example, in U.S. Pat. Nos. 1,692,845 entitled "Internal 
Combustion Engine" issued Nov. 27, 1928 to George A. Kolb; 2,011,864 
entitled "Pump" issued Aug. 20, 1935 to Sverrer H. Lundh; 2,602,434 
entitled "Hydraulic Valve Operating Mechanism Operable to Vary Valve Lift 
and Valve Timing" issued July 8, 1952 to James C. Barnaby, and 3,257,999 
entitled "Hydraulic Control for Internal Combustion Engines, in Particular 
for Gas Engines" issued June 28, 1966 to Franz Fiedler. 
In such systems as disclosed in the above-identified United States patents, 
a control pump driven by the engine is used to intermittently supply high 
pressure fluid via a check valve and a delivery pipe to one side of a 
hydraulic cylinder having a piston reciprocably journaled therein whereby 
the piston can be actuated in one direction to affect opening of a poppet 
valve against the bias of the usual valve spring effecting normal closing 
of the poppet valve. The control pump in such a system is usually a 
conventional type of fuel injection pump, as used, for example, to supply 
pressurized fuel in the fuel injection system of a diesel engine, and, as 
well known, a single control pump may be used to supply the hydraulic 
fluid for actuation of a plurality of poppet valves in a set of valves for 
the engine or, alternatively, individual pumps, of the so-called jerk type 
commonly used to effect fuel injection in diesel engines, may be used for 
each poppet valve of the engine. 
However, none of the known prior art hydraulic valve actuator systems 
provide control of valve lift, valve duration, and valve timing in a 
simple, inexpensive arrangement that can be readily incorporated into an 
engine or similar mechanism. 
SUMMARY OF THE INVENTION 
This invention relates to a hydraulic valve actuator system of the type 
having an engine driven control pump, similar in construction to a 
conventional fuel injection pump, that is used to intermittently supply a 
metered quantity of pressurized hydraulic fluid, as a function of engine 
operation, to a hydraulic valve actuator cylinder having a piston therein 
positioned to abut against the stem of a poppet valve to effect opening 
movement of a poppet valve that is normally held in a closed position by a 
valve return spring. Each delivery conduit connecting an outlet of the 
control pump to a hydraulic valve actuator cylinder is also connected, 
intermediate its ends, to a relatively low pressure fuel drain conduit, 
with flow through the drain conduit being controlled by a normally open, 
solenoid valve connectable to a suitable electrical control circuit which 
is operative to effect sequential energization of that solenoid valve as a 
function of engine operation, as predetermined for a particular engine. 
It is therefore a primary object of this invention to provide a hydraulic 
valve actuator whereby valve lift, timing and valve duration can be 
controlled, as desired, for effective and economic engine operation. 
Another object of this invention is to provide a hydraulic valve actuator 
system that is operable to provide improved control of engine breathing 
over all engine speed and load operating conditions. 
A further object of this invention is to provide a hydraulic valve actuator 
system whereby an internal combustion engine can be operated with wide 
open throttle at the engine carburetor while controlling engine power 
output via inlet valve control by means of the subject system. 
A still further object of this invention is to provide hydraulic valve 
actuator system for an internal combustion engine whereby the engine can 
be operated in a split engine mode by selective deactivation of engine 
valves.

DESCRIPTION OF THE PREFERRED EMBODIMENT 
Referring now to FIG. 1, there is shown schematically a portion of an 
internal combustion engine 1 having, in the construction illustrated, a 
plurality of longitudinally aligned cylinders 2, only one of which is 
shown, each such cylinder 2 having reciprocably disposed therein a piston 
3 operatively connected to the engine crankshaft, not shown. Mounted on 
the cylinder block and enclosing the upper ends of the cylinders 2 is a 
cylinder head 4 which cooperates with the cylinders to define combustion 
chambers 5 and includes passages 6, each of which terminates at one end in 
a port 7 opening into an associated combustion chamber 5. Each such port 7 
is closeable by an outward opening poppet valve 8. 
Each poppet valve 8 which may be either an intake valve or an exhaust 
valve, is of conventional construction and includes a valve head 10 with a 
valve stem 11 extending therefrom that is suitably journaled and guided in 
a valve stem guide bore 12 in the cylinder head 4 whereby the head 10 of 
the poppet valve can be moved into and out of engagement with the valve 
seat encircling and forming part of a port 7. Each such poppet valve 8 is 
normally biased to a closed position relative to its associated port 7 by 
means of a valve return spring 14 bearing upwardly from the cylinder head 
4 against a spring retainer 15 locked, in a well known manner, to the 
valve stem 11 closely adjacent to its free end. 
Suitably supported above each of the poppet valves 8 is a hydraulic 
actuator means that includes an actuator housing 20 which may be formed as 
part of the cylinder head 4 or, as shown, as a separate element. Housing 
20 is provided with hydraulic valve actuator cylinders 21 therein, the 
axis of each such actuator cylinder 21 being aligned with the axis of the 
valve stem 11 of the poppet valve 8 with which it is associated. An 
actuator plunger or piston 22 is reciprocably journaled in each of the 
actuator cylinders 21, whereby each such actuator cylinder 21 and actuator 
piston 22 forms, in effect, a valve actuator. Each actuator piston 22 is 
suitably structured so that one end thereof, its lower end with reference 
to FIG. 1, operatively abuts against the free end of the valve stem 11 
with which it is associated. Thus, in the construction illustrated, each 
actuator piston 22 has a reduced diameter extension 23 at its lower end 
that slidably projects outward through an opening 24 provided in the lower 
end of the actuator housing 20 in axial alignment with the associated 
actuator cylinder 21. Also in the construction illustrated, each actuator 
piston 22 is provided at its opposite or upper end, with reference to FIG. 
1, with a pilot plunger 25 of predetermined diameter and axial length for 
a purpose to be described. 
Each actuator cylinder 21 in the housing 20 is partly closed at its upper 
end, with reference to FIG. 1, by a housing cylinder head 26 having a 
passage 27 therethrough, opening into that actuator cylinder 21. This end 
of each passage 27 opening into an actuator cylinder 21 is of a 
predetermined internal diameter so as to loosely receive the pilot plunger 
of the associated actuator piston 22. The opposite end of each passage 27 
is suitably connected to one end of an associated delivery conduit 28 
whereby the associated actuator cylinder can be intermittently supplied 
with a metered quantity of hydraulic fluid, such as oil, in a manner to be 
described, whereby, as a predetermined hydraulic force is applied against 
the actuator piston 22, it will be moved in a power stroke direction to 
affect opening movement of the poppet valve 8, and when this force is 
reduced sufficiently, the valve return spring 14 will affect reseating or 
closure of the poppet valve 8 and, at the same time, cause movement of the 
actuator piston 22 in a return stroke direction. 
In the embodiment disclosed, pressurized hydraulic fluid is intermittently 
supplied to each actuating cylinder 21 via its associated delivery conduit 
28 by an engine driven distributor pump, generally designated 30, which in 
turn is supplied with hydraulic fluid from a fluid reservoir 31 via a 
supply conduit 32. 
In accordance with the invention, each delivery conduit 28 is connected 
intermediate the one-way check valve 33 controlling flow through this 
conduit and its connection to an associated actuator cylinder 21 by a 
drain conduit 34 to the supply conduit 32 intermediate the ends thereof, 
with flow through the drain conduit 34 to the supply conduit 32 being 
controlled by a normally opened, solenoid valve 35. 
For a purpose which will become apparent, the fluid reservoir 31 should 
either be located in the manner shown to act as an accumulator to provide 
a slightly pressurized column of fluid in the supply conduit 32 so as to 
prevent complete drainage of fluid from the associated delivery conduit 28 
when the associated solenoid valve 35 is open or alternately, the fluid 
reservoir 31 should be in the form of a low pressure type accumulator 
chamber, not shown. As a further alternate, the fluid reservoir 31 can be 
located as desired and a low pressure supply pump, not shown, can be used 
to deliver fluid from the fluid reservoir 31 to the pump 30 through the 
supply conduit 32 at a predetermined low supply pressure in the same 
manner as currently used in certain diesel engine powered vehicles. In 
this second alternate construction, such a supply pump should be located, 
for example, in the supply conduit 32 up stream of the solenoid valve 35 
controlled flow opening from the drain conduit 34 thereto. 
Distributor pump 30 may be any suitable type positive displacement pump 
such as those presently available for use in diesel fuel injection 
systems. Accordingly, the distributor pump 30 may be of the type disclosed 
in U.S. Pat. No. 3,648,673 entitled "Fuel Injection Pump" issued Mar. 14, 
1972 to Richard S. Knape or, as schematically shown, the pump may be of 
the type disclosed, for example, in U.S. Pat. No. 3,861,833 entitled "Fuel 
Injection Pump" issued Jan. 21, 1975 to Daniel Salzgerber, Robert 
Raufeisen and Charles W. Davis. This later type pump is, in effect, an 
engine driven inlet-metering, distributor type pump having a built-in 
governor and automatic advance mechanism whereby the pump is operative to 
provide timed delivery of metered quantities of pressurized fluid as a 
function of engine operation. 
As is known, a pump of the type disclosed in the above identified U.S. Pat. 
No. 3,861,833 includes a pump housing 40 having an inlet 41 connected to 
the common supply conduit 32 through which fluid flows to an internal 
transfer pump 42. Since the displacement of the transfer pump 42 greatly 
exceeds final pump 30 discharge requirements, a large percentage of fluid 
delivered by the transfer pump is by-passed back to the inlet side of the 
transfer pump 42 through a by-pass means having a pressure regulating 
valve 43 therein which causes both the amount of fluid by-passed and 
transfer pump 42 discharge pressure to increase with engine speed. A 
portion of the fluid discharged from the transfer pump is forced through a 
metering valve 44, as regulated by engine demand through a governor 
arrangement, not shown. From the metering valve 44, fluid flows through a 
check valve 45 in a passage 46 to a distributor rotor 47 rotatably 
journaled in housing 40 and driven in a suitable manner, not shown, by 
engine 1, as is the transfer pump 42. With this arrangement, pump 
operation is synchronous with engine 1 operation. When one of the charging 
port passages 48 in the rotor 47 comes into register with the passage 46, 
fluid flows into the injection pump chamber 50 in rotor 47. 
The injection pump chamber 50 is formed by a pair of intersecting 
transverse bores in the rotor 47. A pair of opposed plungers 51 are 
mounted for reciprocating movement in these bores. Surrounding the 
distributor rotor 47 is a generally annular internal cam ring 52, which, 
in turn is journaled in housing 40 for limited arcuate movement and is 
moved by means of an advanced mechanism, not shown, to which it is 
operatively connected by means of a connecting pin 53. The advanced 
mechanism, not shown, is operative whereby delivery of fuel by the 
injection pump from the injection pump chamber 50 will be varied as a 
function of engine operation. Cam rollers 54 and cam roller shoes 55 are 
carried by the rotor 47 between the plungers 51 and the cam ring 52. 
When metered fuel is admitted to the injection pump chamber 50, the 
plungers 51 move radially outward, as required, to receive the charge of 
fluid delivered thereto. At that time, the cam rollers 54 are positioned 
between adjacent cam lobes 52a of the cam ring 52. Rotation of rotor 47 
then causes the charging port passage 48 to pass out of registry with 
passage 46 and, as the distributing passage 56, shown schematically, comes 
into registry sequentially with one of a plurality of passages 57 to the 
outlets 58, cam rollers 54 simultaneously contact opposite lobes 52a of 
the cam ring 52 to effect reciprocable motion of the plungers 51 so as to 
pressurize the charge of fluid in the injection pump chamber 50 on the 
inward stroke of the plungers 51 to some predetermined high pressure. This 
pressurized fluid is, of course, then discharged via one of the passages 
57 to its associated outlet 58 that, in turn, is connected to a delivery 
conduit 28, previously described. It should be noted that only one of the 
passages 57 and only one of the outlets 58 is shown in FIG. 1, it being 
realized that an N number of such passages and an N number of associated 
outlets would be provided on a particular embodiment of the distributor 
pump 30 for a particular engine application having an N number of 
cylinders. 
It is believed that the foregoing description is sufficient for the purpose 
of this application to show the general operation of this particular type 
engine driven distributor pump 30. For further details concerning the 
specific construction of this type distributor pump 30, as shown, 
reference is made to the above-identified U.S. Pat. No. 3,861,833, the 
disclosure of which is incorporated herein by reference thereto. 
For purposes of disclosure only, the engine 1 is referred to as a V-8 
engine. It will thus be apparent that in such an engine there would be two 
banks of cylinders 2 and that the total of N number of cylinders in this 
engine would be eight. It will also be apparent that in such an engine it 
would be provided with at least an N number or eight inlet or intake 
poppet valves and at least an N number or eight outlet or exhaust poppet 
valves for controlling ingress of induction fluid to the cylinders 2 and 
egress of exhaust gas from the cylinder. 
In such an engine, a first system, as shown in FIG. 1, with a first 
distributor pump 30 would be required to operate the N number of inlet 
valves and a second similar system with a second distributor pump 30 would 
be required to operate the N number of exhaust valves. The outlets 58 of 
each of the first and second distributor pumps would, of course, 
correspond in number to the N number of cylinders of the engine, with 
these outlets 58 supplied sequentially with pressurized fluid by the 
rotating distributor rotor 47 with the two plungers 51 (or sometimes four, 
as is well known) therein, as previously described, the plungers being 
driven by the cam lobes 52a, the number of which will also depend on the 
number of outlets in a particular pump, as is well known in the art. The 
distributor rotor 47, as well as the driven element of the transfer pump 
42 of each distributor pump is suitably coupled to the engine 1 via a 
drive system, not shown, so as to operate at half engine crankshaft speed. 
In the subject system, it is preferred that the check valves 33 and 
solenoid valves 35 are located in close proximity to the distributor pump 
30 so that, in effect, a single hydraulic line is required to connect each 
remote valve actuator to the distributor pump 30. This hydraulic line is 
thus operative to accommodate fluid flow in both directions to facilitate 
valve opening and valve closing. 
As previously described, the accumulator or reservoir 31 is operatively 
connected to each pump outlet check valve 33 and associated valve actuator 
by a normally open solenoid valve 35. Each such solenoid valve 35 is a 
normally open, so-called poppet type valve and it is used in such a manner 
so as to utilize the greatest magnetic holding force in the valve closed 
position and so as to release, when de-energized, very quickly due to 
hydraulic pressure attempting to unseat the valve thereof. As will be 
apparent, each solenoid valve 35 would be connected to a suitable source 
of electric power, not shown, through a suitable valve timing electrical 
computer control device, whereby these solenoid valves 35 can be 
sequentially energized in a predetermined manner as a function of engine 
operation and/or operator input, as predetermined. Since the details of a 
suitable valve timing control device and its operation is not deemed 
necessary for an understanding of the subject invention, such a device has 
not been disclosed herein, especially in view of the fact that such 
control devices are well known in the electronic art as associated with 
the fuel injection system, for example, of automotive engines as shown, 
for example, in U.S. Pat. Nos. 3,240,191, Wallis and 3,817,099, Bubniak et 
al, and a description of such a device is not deemed necessary for an 
understanding of the subject invention. 
As previously described, each valve actuator, consists of a hydraulic 
piston 22 that is operative to generate a downward force on an associated 
valve 8 when fluid pressure is applied to the upper end, with reference to 
FIG. 1 of the piston 22. The applied fluid pressure is predetermined and 
is always sufficient to overcome resisting forces of the valve return 
spring 14, friction, and cylinder pressure because of the positive 
displacement nature of the system. The use of a valve actuator in 
conjunction with a conventional valve return spring 14, provides, in 
effect, a hydraulic pump on the return stroke of the valve 8. Thus, 
hydraulic energy is conserved by storing the energy return on the return 
stroke of the valve actuator in the pressure accumulator or reservoir 31. 
It will readily be apparent that the valve return spring 14 should have 
sufficient force to assure that the valve 8 will fully return to its 
closed position and so that it can not be unseated by residual pressure in 
the conduit 28 when the associated solenoid valve 35 is open. The pilot 
plunger 25, previously described, is included on the actuator piston 22 so 
as to orifice the flow of fluid from the actuator cylinder 21 during the 
final portion of the return stroke of the actuator piston 22 therein 
whereby to provide a hydraulic cushion for gentle seating of the valve 8 
against its valve seat and to effect a reduction in noise. 
In the operation of the subject hydraulic valve actuator system, valve lift 
control and valve timing control is achieved by normal operation of the 
engine driven, distributor pump 30, while valve duration control is 
achieved by operation of the solenoid valves 35, as follows: 
Valve Lift Control 
Valve lift control is achieved in the subject system through the operation 
of the metering valve 44 of the distributor pump 30. The quantity of fluid 
admitted by the metering valve during the suction stroke of plungers 51 
determines the amount of valve 8 displacement per stroke in a 1:1 manner. 
The valve 8 lift will thus occur in a manner that maintains the maximum 
lift point in a constant relationship to piston 3 position. Thus, at a 
given time setting, maximum valve lift will always occur at the same point 
(e.g., 90.degree.) regardless of amount of lift as shown in FIG. 2. This 
feature provides inherent advance of valve opening as greater valve lifts 
are generated in response to system controls. 
Valve Timing Control 
Valve timing control in the subject system is achieved by rotation of the 
cam ring 52 with respect to the engine crankshaft in a manner described in 
detail in the above-identified U.S. Pat. No. 3,861,833. Movement of the 
cam ring 52 sets the point when the plungers 51 contact the cam lobes 52a 
at which time lift of an engine valve 8 begins. Although a particular 
mechanism has been shown in the above-identified U.S. Pat. No. 3,861,833 
to effect rotation of the cam ring 52 as a function of engine operation, 
various other types of vacuum or electrical actuators can be used, as 
known in the art, to effect rotation of such a cam ring 52 in this type 
distributor pump 30. 
Valve Duration Control 
Valve duration control is achieved by the associated solenoid valve 35 for 
each valve 8. The electrical control device, not shown, is operative so 
that the solenoid valve 35 must be energized (closed) when its associated 
valve 8 lift begins. When maximum valve lift is attained, the distributor 
pump 30 output check valve 33 locks a holding pressure on the associated 
valve actuator and holds that associated valve 8 open against the force of 
its valve return spring 14. When this solenoid valve 35 is de-energized, 
the hydraulic lock is lifted and then the associated valve return spring 
14 is operative to return that valve 8 to its seated or closed position, 
and at the same time to act against the associated actuator plunger 22 to 
effect a return stroke thereof whereby to force the fluid back to the 
accumulator or reservoir 31. Thus, the valve lift profile follows the 
profile of the cam lobes 52a as seen in FIG. 2 and valve opening time 
become a function of engine speed. In contrast, valve closing time is a 
hydraulic function that remains constant over the engine speed range. 
Valve duration control, as shown in FIG. 2, simply delays hydraulic 
release for the desired time or degrees of crankshaft position. 
Referring now to FIG. 3, there is shown a series of six representations of 
both intake and exhaust valve operation, that can be programmed as 
desired, for an engine equipped with separate hydraulic valve actuator 
systems, in accordance with the invention, for the intake valves and 
exhaust valves of the engine. Line one of FIG. 3 illustrates typical valve 
operation over the full range of speeds and loads for a conventional fixed 
camshaft engine. Lines two thru six of FIG. 3 illustrate a variety of 
operating modes using "Valve Lift, Timing, and Duration Control" by use of 
hydraulic valve actuator systems as disclosed herein. 
Thus, for example, line two of FIG. 3 illustrates part-power operation for 
low to mid speed engine operating conditions. The objective of this type 
valve programming is to eliminate valve overlap and maximize both the 
expansion ratio and the compression ratio of the engine for efficiency and 
economy. 
Line three of FIG. 3 illustrates a maximum power, high speed engine 
operation which is utilizing maximum valve lift as well as advanced 
opening and retarded closing of both intake and exhaust valves. The curves 
in line three are those typical of a high performance engine camshaft. 
Line four of FIG. 3 is another illustration of part-power engine operation 
showing the early closing of the exhaust valve whereby to trap residual 
exhaust gases and thereby accomplish, in effect, internal exhaust gas 
recirculation. 
Line five of FIG. 3 illustrates a low speed, low power, engine operating 
condition where a very limited lift of the intake valve is combined with 
early closing to restrict the intake of a combustion mixture into the 
combustion chamber of the engine. This mode of operation of the intake 
valve accomplishes the throttling of the engine and is operative to reduce 
engine pumping losses because the throttling can be restricted to a very 
brief time period. 
Line six of FIG. 3 illustrates a by-passed cylinder operating in a split 
engine mode, that is, a condition wherein both the intake valve and 
exhaust valve for a particular cylinder are deactivated. Operating the 
engine on reduced cylinders by means of the subject hydraulic valve 
actuator systems complements the function shown in line five of FIG. 3 by 
maintaining the operating points of the remaining operating cylinders in a 
range that is readily controlled by variable valve operation as disclosed 
herein. 
It will be apparent to those skilled in the art that the subject hydraulic 
valve actuator system when used to control operation of the intake valves 
is operative to provide effective intake valve throttling for a sonic 
throttling intake valve engine as disclosed, for example, in U.S. Pat. No. 
3,422,803 entitled "Internal Combustion Engine Construction and Method for 
Operation with Lean Air-Fuel Mixtures" issued Jan. 21, 1969 to Donald L. 
Stivender. Although preferably in such a sonic throttling intake valve 
engine separate systems, in accordance with the invention, would be used 
for both the intake and exhaust valves, if desired, a conventional 
actuating mechanism could be used to effect operation of the exhaust 
valves. 
Referring now to FIG. 4 which is a block diagram of a control system for 
the operation of both the inlet valves and exhaust valves of an engine 
using the normal operating characteristic of a pair of engine driven 
distributing pumps 30 in separate hydraulic valve actuator systems for the 
inlet valves and exhaust valves for controlling the valve lift and timing 
of these valves supplemented by the electronic control of the solenoid 
valves 35 in each system to regulate valve open duration. 
As shown in FIG. 4, an operator actuated accelerator pedal 60 is 
mechanically connected in a known manner to each distributor pump 30 to 
control operation of the metering valve 44 of each pump 30 and, an 
electrical signal is generated proportional to accelerator pedal 60 
position. Since this signal is typically analog and the electronic control 
device 61 is typically a digital electronic processor, an analog/digital 
converter 62 is used to convert the analog signal to a digital signal 
applied to the electronic control device 61. In the control system shown, 
two additional analog inputs are also shown as being multiplexed into the 
analog/digital converter, one such input being applied by a temperature 
sensor 63 and the other input being applied by an altitude pressure sensor 
64. 
Temperature compensation is used to delay inlet valve closing with 
increasing temperature to compensate for reduced density of air at 
elevated temperatures. Altitude compensation is based on an absolute 
pressure transducer that would also delay inlet valve closing for reduced 
pressures (and density) of the induction air. 
An engine revolution per minute (RPM) signal can typically be tapped off 
the usual sensor 65 for the ignition system of the engine, which generates 
four pulses per revolution of a V-8 engine. These pulses would be counted 
in the electronic control device 61 for a preselected time period to 
determine engine speed. The electrical signal inputs of accelerator pedal 
position and RPM would identify a unique value of crank angle for precise 
closing of the valves. When the engine is operating in a split mode, the 
accelerator pedal 60 input would be doubled in the electronic control 
device 61 to determine the proper closing event for the cylinders 
remaining in operation, that is, 30% accelerator pedal movement (all 
cylinders operating) corresponds, for example, to 60% accelerator pedal 
position for half the cylinders in operation. The crank angle to trigger 
valve closing may be based on a proportion of time between distributor 
pulses or it may be based on a toothed sensor of sufficient resolution to 
establish a crank angle for valve closing, that is, a sensor of the type 
used in a conventional electronic spark timing system. The parameters of 
temperature and altitude operate on the selected crank angle by adding 
degrees (delay) to compensate for reduced air density, as previously 
described. 
It will be apparent to those skilled in the art that various modifications 
can be made in the subject hydraulic valve actuator system without 
departing from the scope of the invention disclosed. For example, the 
usual governor control input to the metering valve of the distributor pump 
30 of the type disclosed in the above-identified U.S. Pat. No. 3,861,833 
could be eliminated so that fluid metering is dependent solely on 
accelerator pedal position.