U.S. patent number 3,926,159 [Application Number 05/454,683] was granted by the patent office on 1975-12-16 for high speed engine valve actuator.
Invention is credited to Gunnar P. Michelson, Louis A. Ule.
United States Patent |
3,926,159 |
Michelson , et al. |
December 16, 1975 |
High speed engine valve actuator
Abstract
An electro-hydraulic system capable of high speed and precision
timing, and particularly adapted to the opening and closing of the
intake and exhaust valves of an internal combustion engine under
computer control. The actuator is powered by pressurized crankcase
oil which is discharged on the engine valve stems for lubrication,
employs hydraulic cushioning to reduce engine valve seating
velocity, and recovers energy on the return stroke of the valve to
improve efficiency. High speed is achieved by rapid deenergization
of electromagnets which hold the fluid control valves closed
against the force of the pressure of the hydraulic supply. These
electromagnets latch the valves in their most power-efficient mode,
that is, with the magnetic circuit gap at or near zero.
Inventors: |
Michelson; Gunnar P. (Santa
Barbara, CA), Ule; Louis A. (Rolling Hills, CA) |
Family
ID: |
23805645 |
Appl.
No.: |
05/454,683 |
Filed: |
March 25, 1974 |
Current U.S.
Class: |
123/90.11;
123/90.12; 123/90.13; 123/90.49 |
Current CPC
Class: |
F01L
9/10 (20210101); F01L 9/20 (20210101) |
Current International
Class: |
F01L
9/04 (20060101); F01L 9/02 (20060101); F01L
9/00 (20060101); F01L 009/04 () |
Field of
Search: |
;251/30
;123/90.12,90.11,90.13,90.49,14MC |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Myhre; Charles J.
Assistant Examiner: O'Connor; Daniel J.
Attorney, Agent or Firm: Nardelli; Dominick
Claims
What is claimed is:
1. A control device comprising:
an internal combustion engine having at least one intake valve and
one exhaust valve;
an hydraulic cylinder and piston assembly having a piston slidably
disposed within a cylinder for actuating a respective one of said
intake and exhaust valves;
pump and fluid supply means for supplying pressurized fluid to the
interior of each cylinder of said assemblies so that said piston
slides therein;
each of said assemblies having an inlet valve and an outlet valve,
both disposed to open under the influence of said high pressure
fluid;
a solenoid for each one of said inlet and outlet valves, said
solenoid being disposed to hold the respective inlet or outlet
valve closed while energized;
a computer means responsive to the output of said internal
combustion engine for supplying a first signal to deenergize said
solenoid associated with said respective inlet valve and energizing
said solenoid associated with said respective outlet valve, so that
said respective piston is forced out of said cylinder by said
pressurized fluid and for supplying a second signal to de-energize
said solenoid associated with said respective outlet valve and
energizing said solenoid associated with said inlet valve so that
said piston is free to move into said cylinder.
2. The device of claim 1 wherein:
a variable speed drive is coupled to said engine to transmit energy
from said engine to said pump means.
3. The device of claim 2 wherein:
an hydraulic accumulator is provided between said pump means and
said assemblies to produce controlled pressure surges; and
said fluid is a liquid.
4. The device of claim 3 wherein:
said cylinder has a plurality of fluid outlets axially spaced
therein so that respective ones of said fluid outlets are closed as
said piston moves into said cylinder; and
each of said fluid outlets communicates with said respective outlet
valve.
5. The device of claim 4 wherein said hydraulic cylinder and piston
assembly has:
a first duct wherein said inlet valve is disposed to supply
pressurized fluid into the cylinder thereof;
a second duct coupled between the cylinder thereof and the supply
of pressurized fluid, and
a relief valve disposed within said second duct, disposed to allow
fluid to flow from the cylinder back to the supply means.
6. The device of claim 1 wherein:
an hydraulic accumulator is provided between said pump means and
said assemblies to produce controlled pressure surges; and
said fluid is a liquid.
7. The device of claim 6 wherein:
said cylinder has a plurality of fluid outlets axially spaced
therein so that respective ones of said fluid outlets are closed as
said piston moves into said cylinder; and
each of said fluid outlets communicates with said respective outlet
valve.
8. The device of claim 7 wherein said hydraulic cylinder and piston
assembly has:
a first duct wherein said inlet valve is disposed to supply
pressurized fluid into the cylinder thereof;
a second duct coupled between the cylinder thereof and the supply
of pressurized fluid, and
a relief valve disposed within said second duct, disposed to allow
fluid to flow from the cylinder back to the supply means.
Description
BACKGROUND OF THE INVENTION
1. Field of the Invention
The present invention relates to the art of operating fluid valves,
requiring high power for their actuation, at high speed and with
high precision; in particular, it offers the capability of improved
opening and closing of the intake and exhaust valves of an internal
combustion engine under computer control.
2. Description of Prior Art
A major obstacle to the more efficient operation of internal
combustion engines, and of many other contrivances that employ
fluid valves, is the power required to open and close such valves
at high speed. This power often exceeds the power that can be
developed by a practical solenoid actuator and though hydraulic
actuators controlled by servo-valves are capable of supplying the
necessary speed and force, these latter are both expensive and
excessively inefficient. For nearly a century the intake and
exhaust valves of internal combustion engines have therefore
employed a camshaft and heavy valve springs to provide the high
force required to open and close intake and exhaust valves. This
rigid, inflexible and "brute force" method of valve actuation is
devoid of all possibility of adapting the timing of valve opening
and closing to the variable speed and load of the engine, with the
consequence that engine torque and efficiency are sacrificed at
both low and high speeds. The choice of the cam profile for the
camshaft of an internal combustion engine is a compromise which has
no truly satisfactory solution.
In a previous patent application, "Programmed Valve System for
Internal Combustion Engines," Ser. No. 306,399, filed Nov. 14,
1972, as a continuation-in-part of a prior patent application, Ser.
No. 125520, filed Mar. 14, 1971, (both assigned to the assignee of
the present application), two classes of engine valve actuators
were described, both hydraulically powered; one in which the valve
actuators were controlled by engine driven hydraulic control valves
whose timing was susceptible to computer control, and a second
class in which the engine valve actuators were controlled by a
single stage solenoid hydraulic valve. In that prior patent
application, means were also shown for operating the solenoid at
high speed and high efficiency, using inductors as energy storage
and recovery means. In a subsequent patent application, "High Speed
Electromagnet Control Circuit," Ser. No. 377,956, filed July 10,
1973 as a continuation-in-part of a prior patent application, Ser.
No. 308,268, filed Nov. 21, 1972, an improved circuit was described
for the high speed, high efficiency control of electromagnets by
means of switching circuits only.
Both of the copending aforementioned U.S. patent applications, Ser.
No. 306,399, filed Nov. 14, 1972, and 377,956, filed July 10, 1973,
are incorporated herein by reference.
SUMMARY OF THE INVENTION
The present invention is a hydraulically powered two-position
actuator adapted to the opening and the closing of the intake and
exhaust valves of an internal combustion engine under electronic
control and at high speed and efficiency. Increased speed and
efficiency over previous methods are made possible by reducing the
power required of the solenoid for the actuation of the hydraulic
fluid control valves. The solenoids and associated electromagnets
are not required to produce a high force over the full travel of
the associated hydraulic control valves, but are only required to
produce a high force in the closed position, when the magnetic air
gap is zero or at least small and, correspondingly, the electrical
power required is also small, in which position the magnetic force
holds the hydraulic control valve closed against the force of
hydraulic pressure. When the solenoid is de-energized, the
hydraulic control valve is forced open at high speed by the
pressure of the hydraulic fluid. The inlet and outlet valve to the
cylinder of a hydraulic actuator are alternately held closed
against the supply pressure by their respective solenoids, which
are alternately energized. When the inlet, solenoid-controlled
valve is opened, the hydraulic supply pressure acts upon the
actuator piston (the outlet solenoid-controlled valve meanwhile
being held closed) and drives the actuator piston to its extreme
position, after which time the inlet control valve acts as a check
valve to prevent return of hydraulic fluid to the supply. When the
outlet valve is released, the inlet valve (having already closed)
is held closed against the increasing force acting to open it as
pressure in the cylinder drops. In the absence of magnet force
acting to close an inlet or outlet control valve, these valves are
returned to their closed position by return springs. The
requirement for hydraulic power is reduced by three ancillary
means; first, a buffer which allows pressure to build up from a
constant flow hydraulic supply to provide a high peak pressure for
actuator operation initially, while allowing a lower average
hydraulic supply pressure and power requirement, second, by not
hydraulically driving the engine valve to a stop which absorbs its
kinetic energy, but by using its kinetic energy to compress the
engine valve spring and then to lock it hydraulically when its
speed has dropped to zero, and, third, by recovery of part of the
kinetic energy of the load, (typically, the engine intake or
exhaust valve and associated spring) this kinetic energy being used
to return part of the fluid in the actuator cylinder to the
supply.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a block diagram of an internal combustion engine in which
the intake and exhaust valves are operated by hydraulically powered
actuators under computer control, according to the invention.
FIG. 2 is a semi-schematic, cross-sectional view of a typical
actuator of the invention mounted over a valve of an internal
combustion engine.
FIG. 3 is a cross-sectional view of the hydraulic buffer.
FIG. 4 is a cross-sectional view of the actuator incorporating a
means for energy recovery.
DESCRIPTION OF THE PREFERRED EMBODIMENT
Unlike a conventional camshaft, used to open and close the intake
and exhaust valves of an internal combustion engine,
electrohydraulically controlled engine valve actuators require a
separate timing circuit or computer to determine the optimum
opening and closing times under varying conditions of engine
operation, such as low speed or high speed forward or reverse,
acceleration or deceleration. A typical computer-controlled valve
actuating system is shown in FIG. 1 where computer 1 receives a
variety of input data from the human operator 3, the environment 5,
and the engine 2, which data are used to compute timing signals 29,
30, 66 and 67 for the valve actuators 16, as well as control
signals for other functions not shown in FIG. 1, such as ignition
and fuel metering. Operator data include such commands as start,
accelerate, decelerate and reverse, such data being communicated to
the computer by transducers which are the equivalent of the
accelerator pedal and other conventional controls. Environmental
data are principally air temperature and air density. Data that the
computer requires of the engine are instantaneous crank angle 4,
intake manifold temperature 6 and possibly other data such as onset
of detonation. Typical computer outputs required for the proper
operation of the engine are fuel flow rate 7, ignition timing 8,
intake valve open commands 9, intake valve close commands 11,
exhaust valve open commands 12, exhaust valve close commands 10,
and the desired valve lift 13. Each intake valve (shown typically
as 14) or each exhaust valve (shown typically as 15) is controlled
by a hydraulic actuator 16. The computer design, per se, does not
form a part of the present invention, but is well documented in the
aforementioned U.S. patent application Ser. No. 306,399.
Description of such matters as contained herein are for the purpose
of establishing an environment for the present invention and to
indicate its particular utility in an illustrative system. Each
inlet control valve 18 is under the control of a solenoid 19. In
like fashion each outlet control valve 20 is under the control of a
solenoid 21. A pair of solenoids 19 and 21 are alternately
energized by outputs 9 and 11 respectively from an electronic drive
circuit 28 responsive to computer commands 29 and 30 for the
opening and closing of the engine intake valves, and a like pair of
solenoids 19 and 21 are alternately energized by outputs 12 and 10
respectively from an electronic drive circuit 65 responsive to
computer commands 66 and 67 for the opening and closing of the
engine exhaust valves.
The hydraulic actuators 16 are powered by high-pressure oil,
pressurized and moved by the fixed displacement oil pump 23,
drawing engine lubricating oil from the oil sump or crankcase 24 of
the engine 2. The oil drawn from sump 24 is normally filtered by a
conventional filter screen not shown to prevent damage to the pump
23 and is further filtered downstream of the pump 23 by
high-pressure filter 25. Since the pump 23 is driven by the engine
by means of a variable-speed drive 17, the pressure adapts itself
automatically to the current requirement for valve lift, pressure
relief valve 68 acting to prevent excessive pressure surges and the
hydraulic buffer 27 acting to produce desireable peak pressures for
the initial opening of engine exhaust valves particularly and yet
allow a substantially lower average hydraulic pressure. Since the
required hydraulic fluid flow rate is generally proportional to
engine speed, the setting of variable speed drive is constant for a
constant valve lift, regardless of engine rpm. However, it is
desireable to reduce valve lift at low engine speeds, particularly
for the engine intake valves, this in order to produce a high inlet
velocity for the fuel-air charge, needed to adequately mix the fuel
and air for reliable and complete combustion. The reduced valve
lift at low engine speeds is produced by an appropriate setting of
the variable speed pump drive 17 in response to computer command 13
which reduces the flow of hydraulic fluid, viz, engine lubricating
oil, to the actuators 16. The reduced valve lift at low engine
speeds not only improves engine efficiency but also reduces the
hydraulic power requirement and the parasitic load on the engine 2
required to drive the pump 23. Some reduced valve lift at low
speeds is obtained, even without a variable speed drive 17, as the
natural consequence of the reduced low speed volumetric efficiency
of most gear or rotary vane pumps, a factor that can be taken into
account in computing the control parameter 13 for the variable
speed drive 17.
The relative position of the actuator 16 and an engine intake or
exhaust valve is shown in FIG. 2. High pressure engine lubricating
oil, used as a hydraulic fluid, enters passage 37 and is eventually
discharged at relatively high velocity from the outlet control
valve 20, situated above and adjacent to the engine intake or
exhaust valve stem 50. This rather violent discharge of lubricating
oil from the actuator 16 provides ample lubrication of engine valve
stem guides and replaces the conventional means of lubrication
based on delivery of oil to the guides from oil galleries supplying
the tappets via hollow push rods. The components of the engine
intake and exhaust valve mechanisms are otherwise conventional and
include the cylinder head 31, an intake or exhaust valve 15, a
valve stem 50, a spring retainer 32 secured to the valve stem 50,
and the valve spring 33 which holds the valve 15 in its closed
position. The hydraulic actuator 16 is mounted to the cylinder head
31 by means of a structure 69 which may be a portion of the
actuator housing 34.
The hydraulic actuator 16 is comprised of a housing 34, a hydraulic
cylinder 35, a piston 36, a high-pressure oil inlet passage 37, the
oil inlet control valve 18, the oil outlet control valve 20, the
control valve return springs 39 and 40, and the control valve
holding solenoids 19 and 21.
In typical operation, when the engine valve 15 is closed, inlet
control valve 18 will be held closed against the force of the
hydraulic supply pressure by the currently energized holding
solenoid 19; meanwhile, solenoid 21 is de-energized, allowing
outlet control valve 20 to open to relieve any incidental pressure
buildup in the hydraulic cylinder 35 that might be caused by
leakage of inlet control valve 18. Conversely, when the engine
valve 15 is open, outlet control valve 20 is held closed by holding
solenoid 21, now energized while solenoid 19, now de-energized,
allows the inlet control valve 18 to open to allow the full supply
pressure to act upon the cylinder 36 and to constantly supply any
leakage through outlet valve 20. Due to the inertia of the engine
valve 15 it will, after being accelerated to a high speed, compress
the valve spring 33 to a greater degree than possible by static
supply pressure alone acting on the piston 39. When the engine
valve 15 and the piston 36 approach a standstill, the relatively
slight force of return spring 39 will close the inlet control valve
18 which prevents the return of hydraulic fluid to the supply.
Typically, in this condition, the hydraulic fluid trapped in
cylinder 35 will attain a pressure nearly twice that of the supply
pressure, corresponding to a valve lift nearly twice that
obtainable from the average supply pressure only. Operation of
inlet valve 18 as a check valve in the manner just described, in
addition to providing double valve lift, suppresses the oscillatory
motion of the valve 15 that would otherwise ensue.
The holding solenoids 19 and 21 are energized by current flowing
through windings 44 and 45 respectively, which current creates a
magnetic flux that produces a strong attractive force between
armature 41 and stator 42 for solenoid 19, and between armature 49
and stator 43 for solenoid 21. Since the armatures 41 and 49 are
mechanically secured to the associated control valves 18 and 20,
these valves are held in their closed positions by the holding
currents in windings 44 and 45, respectively.
Opening of an engine valve 15 is achieved by the mechanical force
of piston 36 mounted directly over the associated engine valve stem
50, this mechanical force being opposed both by the inertia of the
valve and the force of the valve spring 33. The engine valve 15,
initially accelerated by the superior force of piston 36,
eventually decelerates as the force of the compressed spring comes
to exceed the force of the piston 36, coming to rest at the point
of maximum lift in consequence of both inlet and outlet control
valves 18 and 20 being held closed, the former as check valve in
response to the higher hydraulic pressure in cylinder 35, and the
latter by energized solenoid 21.
The closing of the engine valve 15 is effected by the
deenergization of solenoid 21 and the energization of solenoid 19.
In either case very little work is done by the respective solenoid
since the inlet control valve 18 si closed prior to energizing
solenoid 19, and the outlet valve 20 is opened by the force of
hydraulic pressure in cylinder 35 and connected passage 38, once
the holding force of solenoid 21 is removed. This hydraulic
pressure in cylinder 35 is the result of the force of the engine
valve spring 33 acting on piston 36 via the valve spring retainer
32 and the valve stem 50. At the instant that the outlet control
valve 20 is opened, the hydraulic pressure in cylinder 35 drops
abruptly and the engine valve 15 and piston 36 are accelerated by
the force of the engine valve spring 32 to their closed positions.
Unless means are used to prevent it, the closing speed of engine
valve 15 would be excessively high. Therefore piston 36 closes off
the main outlet passage 38 to gradually produce a braking force
which acts on the said piston and which is opposed to the force of
the engine valve spring 33. When the hydraulic braking force
becomes equal to the spring force, the engine valve 15 and the
piston 36 are at their maximum speed and thereafter their speed
decreases since the hydraulic braking force then exceeds the spring
force. When the piston 36 has completely blocked the large outlet
passage 38, the pressure in cylinder 35 will have reached a rather
high value, since the oil can now escape only through the
throttling passages 46 and the bleeder passage 47. The number,
diameter and the location of the throttling passages 46 are
selected to provide a nearly constant deceleration to an acceptably
low seating velocity for engine valve 15. After the engine valve 15
is seated in its closed position, only the very small bleeder
passage 47 remains open for the purpose of assuring that piston 36
cannot come to a full stop prior to the seating of valve 15 in its
closed position.
When the piston 36 has come to a stop, the flow through the outlet
control valve 20 which had held said valve open has ceased,
enabling the relatively weak force of return spring 40 to close the
outlet control valve 20. Thus the work required to close this valve
20 is supplied indirectly by the hydraulic supply pressure by
energy stored in the return spring 40 so that subsequent
energization of solenoid 21 is not required to produce mechanical
work but only a holding force sufficient to hold outlet control
valve 20 closed against the hydraulic pressure. Although this
holding force is far in excess of that which can be supplied by
return spring 40, it requires relatively little electrical energy.
Thus, with outlet control valve 20 held closed only by the return
spring 40 to prevent pressure build up in cylinder 35 from any
leakage from inlet control valve 18, the actuator is in a state
ready for the near instantaneous opening of the engine valve 15 on
the next cycle, effected by the de-energization of solenoid 19 and
the energization of solenoid 21. To avoid engagement shock between
actuator piston 36 and engine valve stem 50 when the piston is
actuated, a light spring or wave washer 48 is placed between the
shoulder of the piston 36 and the housing 34.
Since the engine valve actuator as described is capable of
producing a mechanical response within a very short period of time,
that is, in the order of 1 millisecond or less from the time of
receipt of computer commands 29, 30, 66 or 67 (FIG. 1.), the time
required to fully open and close an engine valve is determined
essentially by the resonant period of the engine valve 15 and valve
spring 33 combination. In addition to the mass of engine valve 15
there are contributions to inertia from the valve spring 33 and the
piston 36. The opening time of the engine valve does not depend on
the speed of the engine nor on the amount of the valve lift if a
single linear valve spring 33 is used. The time for the valve to
reach maximum lift will be one-half the resonant period of the
mass-spring combination. The time to close, after the opening of
outlet control valve 20, will be one-fourth of this period plus the
time required to decelerate the piston, so that the time to fully
open and subsequently close the engine valve 15 is between
three-fourths and approximately 90percent of the said resonant
period at least, and may be made longer by delaying the "close"
command. Since the resonant period of a typical
valve-mass/valve-spring combination is about 8 milliseconds, full
open to full close times as short as approximately 6 to 7
milliseconds are attainable with conventional linear valve springs.
This minimum open-to-close time, which is fully adequate for an
internal combustion engine, can be made even shorter by the use of
dual valve springs, a slight spring which allows a higher opening
force and achieves a quicker opening of the valve, and a second,
much stiffer spring which engages after sufficient valve lift has
been attained and acts to quickly arrest valve motion and also to
store energy for the rapid acceleration of the engine valve 15 to
its subsequent closed position. Since valve open-to-close time is
dependent on the valve-mass/valve-spring resonant period, shorter
open-to-close times can be attained by the use of stiffer springs
and a corresponding increase in hydraulic power. Further, the time
that an engine valve is open can be reduced by commanding it to
close before it is fully open.
The hydraulic power required to operate the engine valves in the
manner described is proportional to the square of the engine valve
lift, the force of the engine valve spring, and the speed of the
engine. The engine valve lift that will be attained and the
resulting spring force depend on the setting of the variable speed
drive 17. The valve lift is reduced at low engine speeds by means
of the variable speed pump drive 17, which in turn, causes the
hydraulic pressure to drop. Typically, it requires about 5
foot-pounds of work to provide full valve lift but only
approximately one-third this amount for one-half full lift, so that
at low engine speeds a substantial saving in hydraulic power is
possible.
In addition to the average power requirement for engine valve
operation there is the need for a higher than average hydraulic
pressure to open the exhaust valves against both the preload of the
exhaust valve spring and the pressure of combustion products in the
firing chamber. Pressure surges in the hydraulic supply suitable
for the purpose are produced by the combination of a constant flow
hydraulic pump and intermittent flow through the actuators 16 but,
unless moderated, these pressure surges can be excessive. A
properly tuned hydraulic buffer 23 serves both to smooth the
intermittent flow to the actuators to a nearly constant flow
supplied by the pump and to allow the pump output pressure to rise
in the period between actuator cycles to a value that is sufficient
to crack open the exhaust valves but not otherwise excessive. As
typically, in a multi-cylinder internal combustion engine, the
exhaust valve of one cylinder will open simultaneously with the
opening of the intake valves of another cylinder, yet always so
that the exhaust valves are opened first (a greater advance in the
opening of exhaust valves being desireable), the full surge
pressure will always be available to open an exhaust valve and a
slightly depleted surge pressure will be more than adequate to open
the intake valves immediately after. The tuning of the buffer to
achieve the desired pressure surges is effected by a proper choice
of spring constant for spring 53, taking into account the effect of
any mechanical compliance in the hydraulic plumbing. The higher
pressure required to open engine exhaust valves is only momentary,
for once they are cracked open, any pressure in the firing chamber
drops abruptly. A simple version of the buffer 27 is shown in FIG.
3 and consists of a hydraulic cylinder 51, a piston 52, the spring
53, and an adjusting screw 54 for the spring 53. The port 55 of the
cylinder 51 is communicatively connected to the inlet control
valves 19 of the actuators 16 and to outlet port of the high
pressure oil filter 25 supplying oil from pump 23. Because of the
finite spring rate of the spring 53, the oil pressure fluctuates
periodically at a frequency equal to the frequency of operation of
the engine valves. For example, when no actuator requires hydraulic
fluid, the oil delivered by the oil pump 23 enters the buffer 27
through the port 55 and displaces piston 52 against the force of
spring 53 and, as a consequence, the oil pressure rises. The rise
in oil pressure is proportional to the displacement of piston 52
and the spring rate of spring 53. When, during this process, one of
the actuator control valves 18 opens, the actuator 16 will require
a larger oil flow than that provided instantaneously by pump 23,
and this required additional flow is supplied by buffer 27, its
piston 52 returning to its initial position. In this manner a
synchronized pressure fluctuation is obtained, producing pressure
surges, when they are required, with a substantially lower average
hydraulic supply pressure. A further favorable consequence of this
buffered operation is that the initial acceleration of the engine
valve will be higher, producing thereby an opening time less than
one-half the resonant period as previously described.
The buffer spring 53 (FIG. 3.) is selected so that its force range
is adequate to meet the requirements for a fluctuating pressure at
all pressure levels set by the computer 1 acting through the medium
of variable speed drive 17 by command 13 (FIG. 1.). At low pressure
the piston 52 lies close to the bottom 56 of the buffer cylinder
51, and, at higher pressure, it operates a greater distance
away.
The function of the buffer 27, just described, may be provided by
any other suitable known type of hydraulic accumulator having the
appropriate spring rate, such as a container partially filled with
compressed gases. Since the buffer is a principal supplier of
hydraulic fluid during the time of valve actuation, each actuator
may be provided with is own buffer to minimize the hydraulic fluid
supply delay occasioned by the inertia of fluid in relatively long
supply lines. Further, a single buffer may be advantageously placed
downstream of the actuators 16 so that said actuators may be
supplied with hydraulic fluid from two directions for more rapid
response.
Since, typically, the pressure in an energized actuator cylinder 35
will be nearly twice the average hydraulic supply pressure, it
would, in principal, be possible to return a significant portion of
this fluid to the supply with a consequent saving in hydraulic
power. such energy recovery can be achieved by replacing the
throttling passage 38 (FIG. 2.) with a relief valve between the
actuator cylinder 35 and the inlet port 37, as shown in FIG. 4,
where 57 is the added relief valve. When the engine valve 15 is
open, the bottom 58 (FIG. 4.) of the actuator piston 36 will be in
the position depicted by the phantom line 59. The instant the
outlet control valve 20 opens, the piston 36 will move upward, as
described above, except that when it blocks passage 38, most of the
fluid will now be forced through the relief valve 57 back into the
high pressure line 37 and be available to power the next actuator
cycle. A small portion of the hydraulic fluid will escape through
the bleeder 47 which remains open at all times to insure proper
seating of engine valve 15 as previously described. The force of
the relief valve spring 60 is selected to prevent premature opening
of valve 57 during the time that engine valve 15 is being held in
its open position by the hydraulic pressure in chamber 35, since
this latter pressure is substantially higher than the hydraulic
supply pressure in inlet port 37. The optimum distance 64 between
the closing edge 62 of the outlet passage 38 and the fixed locus
defined by the edge 63 of piston 36 in its retracted position, as
also the setting of the adjusting screw 61 for the force supplied
by spring 60, depend on the pressure in the inlet passage 37 and
therefore the energy recovery system just described when equipped
with fixed settings will be most effective for a single fixed
engine valve lift. At larger engine valve lifts the valve seating
will become hard since the kinetic energy of the valve is higher
than the returned energy. At smaller engine valve lifts the valve
will reach a slow speed before it is seated and will take longer to
close since thereafter oil can be discharged only throught the
bleeder passage 47. Therefore, the energy recovery system described
is effective only when a fixed or nearly constant valve lift is
employed.
Modifications and variations within the scope of the invention will
suggest themselves to those skilled in this art once the concepts
of the invention are understood. For example, the foregoing system
for operating the intake and exhaust valves of an internal
combustion engine under computer control using engine lubricating
oil as the hydraulic fluid may be replaced by a system having a
separate hydraulic supply and fluid reservoir, and conventional
means for lubricating valve stem guides may be used in that event.
As another example, the combination of a variable speed drive and a
positive displacement pump may be replaced by a controllable
displacement variable displacement pump.
* * * * *