Pneumatic actuator with permanent magnet control valve latching

Abstract: A bistable electronically controlled pneumatically powered transducer for use, for example, as a valve mechanism actuator in an internal combustion engine is disclosed. The transducer has an armature (13) including a piston which is coupled to an engine valve, for example. The piston is powered by a pneumatic source and includes pneumatic damping with a one-way return of air compressed beyond source pressure back to the air source as it nears its destination position. Air supplied to power the piston may be differentially controlled to compensate for asymmetric resistance to movement and the air damping may be differentially controlled to provide dissimilar damping at the two extremes of piston motion. The armature is held in each of its extreme positions by pneumatic pressure under the control of control valves (15,17) which are in turn held in their closed positions by permanent magnet latching arrangements (21, 23) and are released therefrom to supply air to the piston to be pneumatically driven to the other extreme position by an electromagnetic arrangement (25,27) which temporarily neutralizes the permanent magnetic field of the latching arrangement (21,23).

Description: The present invention is dealing with a pneumatically powered valve actuator comprising a valve actuator housing, a piston reciprocable within the housing along an axis, the piston having a pair of oppositely facing primary working surfaces, a pressurized air source, a pair of air control valves movable between open and closed positions, means for selectively opening one of said air control valves to supply pressurized air from the air source to one of said primary working surfaces causing the piston to move and pneumatic means for decelerating the piston near the extremities of its reciprocation. This actuator finds particular utility in opening and closing the gas exchange, i.e., intake or exhaust, valves of an otherwise conventional internal combustion engine. Due to its fast acting trait, the valves may be moved between full open and full closed positions almost immediately rather than gradually as is characteristic of cam actuated valves. The actuator mechanism may find numerous other applications such as in compressor valving and valving in other hydraulic or pneumatic devices, or as a fast acting control valve for fluidic actuators or mechanical actuators where fast controlled action is required such as moving items in a production line environment. Internal combustion engine valves are almost universally of a poppet type which are spring loaded toward a valve-closed position and opened against that spring bias by a cam on a rotating cam shaft with the cam shaft being synchronized with the engine crankshaft to achieve opening and closing at fixed preferred times in the engine cycle. This feed timing is a compromise between the timing best suited for high engine speed and the timing best suited to lower speeds or engine idling speed. The prior art has recognized numerous advantages which might be achieved by replacing such cam actuated valve arrangements with other types of valve opening mechanism which could be controlled in their opening and closing as a function of engine speed as well as engine crankshaft angular position to other engine parameters. In copending application EP-A-0 281 192 (state of the art according to Art. 54.3 EPC) there is disclosed a valve actuator which has permanent magnet latching at the open and closed positions. Electromagnetic repulsion may be employed to cause the valve to move from one position to the other. Several damping and energy recovery schemes are also included. In copending application EP-A-0 328 195 (state of the art according to Art. 54.3 EPC) there is disclosed a somewhat similar valve actuating device which employs a release type mechanism rather than a repulsion scheme as in the previously identified copending application. The disclosed device in this application is a jointly pneumatically and electromagnetically powered valve with high pressure air supply and control valving to use the air for both damping and as one motive force. The magnetic motive force is supplied from the magnetic latch opposite the one being released and this magnetic force attracts an armature of the device so long as the magnetic field of the first latch is in its reduced state. As the armature closes on the opposite latch, the magnetic attraction increases and overpowers that of the first latch regardless of whether it remains in the reduced state or not. This copending application also discloses different operating modes including delayed intake valve closure and a six stroke cycle mode of operation. In copending application EP-A-0 328 193 (state of the art according to Art. 54.3 EPC) there is disclosed a valve actuating device generally similar in overall operation to the present invention. One feature of this application is that control valves and latching plates have been separated from the primary working piston to provide both lower latching forces and reduced mass resulting in faster operating speeds. This high speed of operation results in a somewhat energy inefficient device. The present application and copending application EP-A-0 347 977 filed on even date herewith address, among other things, improvements in operating efficiency over the above noted devices. Other related applications are EP-A-0 328 194 (state of the art according to Art. 54.3 EPC), where energy is stored from one valve motion to power the next, and EP-A-0 328 192 (state of the art according to Art. 54.3 EPC), wherein in spring (or pneumatic equivalent) functions both as a damping device and as an energy storage device ready to supply part of the accelerating force to aid the next transition from one position to the other. The entire disclosures of all five of these copending applications are specifically incorporated herein by reference. In the present invention, like EP-A-0 328 193, the power or working piston which moves the engine valve between open and closed positions is separated from the latching components and certain control valving structures so that the mass to be moved is materially reduced allowing very rapid operation. Latching and release forces are also reduced. Those valving components which have been separated from the main piston need not travel the full length of the piston stroke, leading to some improvement in efficiency. It is pointed out that a hydraulically driven valve actuator, wherein a working piston, which moves an engine valve between open and closed positions, is separated from a single control sleeve, is known per se from US-A-3,844,528. A pneumatically powered valve actuator according to the opening paragraph is known from DE-C-421 002. In this known actuator the pneumatic means for decelerating the piston is not adjustable and damping air is not vented or recovered. It is the object of the present invention to provide a pneumatically powered valve actuator which is characterized by an improved efficiency, which has improved damping features and which is tolerant of variations in air pressure and other operating parameters. According to the invention this object is essentially obtained in that the pneumatic means includes a one-way pressure relief valving arrangement for venting air from the pneumatic means to the pressurized air source. During movement of the piston from one position to the other air is compressed by the piston and applies an opposing force on the piston to slow piston motion as the piston nears one of the extreme positions. As the piston slows, pressure builds up and when pressure reaches the source pressure, the one-way pressure relief valving arrangement releases part of the compressed air back to the high pressure air source. This as well as other preferred embodiments are defined in the claims and are pointed out hereafter. BRIEF DESCRIPTION OF THE DRAWING Figure 1 is a view in cross-section showing the pneumatically powered actuator of the present invention with the power piston latched in its leftmost position as it would normally be when the corresponding engine valve is closed; Figures 2-9 are views in cross-section similar to Figure 1, but illustrating component motion and function as the piston progresses rightwardly to its extreme rightward or valve open position; and Figures 10 and 11 are views similar to Figure 1, but illustrating certain modifications of the actuator. Corresponding reference characters indicate corresponding parts throughout the several views of the drawing. The exemplifications set out herein illustrate a preferred embodiment of the invention in one form thereof and such exemplifications are not to be construed as limiting the scope of the disclosure or the scope of the invention in any manner. DESCRIPTION OF THE PREFERRED EMBODIMENT The valve actuator is illustrated sequentially in Figures 1-9 to illustrate various component locations and functions in moving a poppet valve or other component (not shown) from a closed to an open position. Motion in the opposite direction will be clearly understood from the symmetry of the components. The actuator includes a shaft or stem 11 which may form a part of or connect to an internal combustion engine poppet valve. The acuator also includes a low mass reciprocable piston 13, and a pair of reciprocating or sliding control valve members 15 and 17 enclosed within a housing 19. The control valve members 15 and 17 are latched in one position by permanent magnets 21 and 23 and may be dislodged from their respective latched positions by energization of coils 25 and 27. The control valve members or shuttle valves 15 and 17 cooperate with both the piston 13 and the housing 19 to achieve the various porting functions during operation. The housing 19 has a high pressure inlet port 39, a low pressure outlet port 41 and an intermediate pressure port 43. The low pressure may be about atmospheric pressure while the intermediate pressure is about 10 psi. above atmospheric pressure and the high pressure is on the order of 100 psi. gauge pressure. Figure 1 shows an initial state with piston 13 in the extreme leftward position and with the air control valve 15 latched closed. In this state, the annular abutment end surface 29 is inserted into an annular slot in the housing 19 and seals against an o-ring 31. This seals the pressure in cavity 33 and prevents the application of any moving force to the main piston 13. In this position, the main position 13 is being urged to the left (latched) by the pressure in cavity or chamber 35 which is greater than the pressure in chamber or cavity 37. In the position illustrated, annular opening 45 is in its final open position after having rapidly released compressed air from cavity 37 at the end of a previous leftward piston stroke. When current flows in coil 25, the field of permanent magnet 21 is partially neutralized and source air pressure on face 49 forces the shuttle or control valve 15 leftwardly against the bias of wave washer 16. In Figure 2, the shuttle valve 15 has moved toward the left, for example, 1.27 mm (0.05 in). while piston 13 has not yet moved toward the right. The air valve 15 has opened because of an electrical pulse applied to coil 25 which has temporarily neutralized the holding force on iron armature or plate 47 by permanent magnet 21. When that holding force is temporarily neutralized, air pressure in cavity 33 which is applied to the air pressure responsive annular face 49 of valve 15 causes the valve to open. Notice that unlike the abovementioned EP-A-0 328 193 application, the communication between cavity 51 and the low pressure outlet port 41 has not been interrupted by movement of the valve 15. This communication is maintained at all times by way of a series of openings such as 54 in control valve 15. It should also be noted that, before the valve clears the slot containing o-ring 31, the edge of air valve 15 has overlapped the piston 13 at 53 closing annular opening 45 of Figure 1 creating a closed chamber to assure rapid pressurization and maximum acceleration of the piston 13. Figure 3 shows the opening of the air valve 15 to about 2.54 mm (0.10 in). (2/3 of its total travel) and movement of the piston 13 about 0.63 mm (0.025 in). to the right. In Figure 3, the high pressure air had been supplied to the cavity 37 and to the face 38 of piston 13 driving that piston toward the right. That high pressure air supply by way of cavity 37 to piston face 38 is cut off in Figure 4 by the edge of piston 13 passing the annular abutment 55 of the housing 19. Piston 13 continues to accelerate, however, due to the expansion energy of the high pressure air in cavity 37. The right edge of piston 13 is about to cut off communication at 57 between the port 43 and chamber 35. Disk 47 is nearing the leftward extreme of its travel and is compressing air in the gap 61. Air control valve 15 has also compressed the wave washer 16. This offers a damping or slowing effort to reduce the end approach velocity and consequently reduce any impact of the air valve components with the stationay structure. The compression of wave washer 16 also stores potential energy to power the return of the control valve 15 to the closed position. The annular surface 62 which is shown as a portion of a right circular cylinder may be undercut (concave) or tapered (a conical surface) to restrict air flow more near one or both extremes of the travel of plate 47 to enhance damping without restricting motion intermediate the ends if desired. The piston 13 is continuing to accelerate toward the right in Figure 4 and the air valve 15 has nearly reached its maximum leftward open displacement. The valve will tend to remain in this position for a short time due to the continuing air pressure on the annular surface 49 from high pressure source 39. There is a bleeding of air between the annular air valve and the piston into chamber 63 which is decreasing the pressure differential across the air valve 15 and this will soon allow the magnetic attraction of the disk 47 by the permanent magnet 21 along with the restorative force from wave washer 16 to pull the air valve 15 back toward its closed position. The wave washer or spring 16 functions as a spring bias means to provide damping of air control valve motion as the air control valve approaches an open position and provides a restorative force to aid rapid return of the air control valve to a closed position. This air bleeding is complete and the motion apparent in Figure 6. In the transition from Figure 4 to Figure 5, the main piston 13 has just closed off communication between chamber 35 and medium pressure port 43 and further rightward motion of the main piston will compress the air trapped in chamber 35 so that the piston will be slowed and stopped by the time it has reached its extreme right hand position. In Figure 5, the air valve 15 is still in its extreme leftward position. The air valve is designed to close at about the same time as the main piston arrives at its furthest right hand location. Also, in Figure 5, the piston is continuing to compress the air in cavity 35 slowing its motion. In Figure 6, the air valve 15 is beginning to return to its closed position. The attractive force of the magnet 21 on the disk 47 and the force of wave washer 16 is causing the disk to move back toward the magnetic latch. Further rightward movement of the piston as depicted in Figure 6, uncovers the partial annular slot 67 leading to intermediate pressure port 43 so that the high pressure air in chamber 36 has blown down to the intermediate pressure. In Figures 6 and 7, the continued piston motion and corresponding buildup of pressure in cavity 35 may cause the pressure in cavity 35 to exceed the source pressure in cavity 33. When this happens, reed valve 101 opens to vent this high pressure air back to the source by way of cavity 33. The reed valves 101 and 103 function to recapture part of the kinetic energy of the piston 13 when damping the piston motion by returning high pressure air to the source 33 rather than merely compressing air in the piston motion damping chamber 35 and then dumping that air to the atmosphere or to the intermediate pressure source. In Figure 7, the pressure in chamber 35 is at its maximum as set by the reed valve 101 and the annular opening is just beginning to form at 69 between the abutting corners of the piston 13 and air valve 17. This annular opening vents the high pressure air from chamber 35 just as the piston nears its right hand resting position to help prevent any rebound of the piston back toward the left. It will be understood from the symmetry of the valve actuator that the behaviour of the air control valves 15 and 17 in this venting or blow-down is, as are many of the other features such as the opening of reed valves 101 and 103, substantially the same near each of the opposite extremes of the piston travel. In each case, the air control valve, piston and a fixed portion of the housing cooperate to vent the damping air from the piston at the last possible moment and after any pressure exceeding that in chamber 33 has been recaptured while these same components cooperate at the beginning of a stroke to supply air to power the piston for a much longer portion of the stroke. The damping of the piston motion near its right extremity is adjustable by controlling the intermediate pressure level at port 43 to effectively control the density of the air initially entrapped in chamber 35. If this intermediate pressure is too high, the piston will rebound due to the high pressure of the compressed air in chamber 35. If this pressure is too low, the piston will approach its end position too fast and may mechanically rebound due to metallic deflection or mechanical spring back. With the correct pressure, the piston will gently come to rest in its right hand position. A further final damping of piston motion may be provided during the last few thousandths of an inch of travel by a small hydraulic damper including a fluid medium filled cavity 73 and a small piston 75 fastened to and moving with the main piston 13. Near either end of the main piston travel, the small piston 75 enters a shallow annular restricted area 77 displacing the fluid therefrom and bringing the main piston to rest. Fluid, such as oil, may be supplied to the damping cavity 73 by way of inlet 85. In Figure 8, the air valve 15 is about midway along its return to its closed position. Final damping is almost complete as the pressure in chamber 35 is being relieved through the annular opening at 69 and through the opening 81 and channel 83 to the low pressure port 41 so that the pressure throughout chamber 35 is reduced to nearly atmospheric pressure. Note that valves 15 and 17 include a number of apertures such as 54 and 81 in their respective web portions allowing free air flow between chambers such as 35 and 83. In Figure 8, the piston 13 is reaching a very low velocity, the damping is almost complete and the final damping by the small fluid piston 75 is underway. The main piston 13 has reached its righthand extreme in Figure 9 and air valve 15 has closed. The supply of high pressure air from the source 39 to chamber 37 and the surface 38 of piston 13 has long since been interrupted by piston edge 105 passing housing edge 55. The piston 13 is held or latched in the position shown by the intermediate pressure in chamber 37 from source 43 acting on piston face 38. In Figure 1, which corresponds to a valve-closed condition, there is a slight gap between the piston face 38 and the valve housing while in Figure 9 with the valve open, no such gap is seen. This gap provides for somewhat greater potential travel of the piston 13 than needed to close the engine valve insuring complete closure despite differential temperature expansions and similar problems which might otherwise result in the engine valve not completely closing. It should also be noted in following the sequence of Figures 1-9 that due to the length of the annular valving surface 107 of piston 13 between the edges 105 and 109, the chamber 63 is never in communication with the high pressure source chamber 33. Chamber 63 is maintained at the outlet pressure of port 41 at all times contrary to the similar chamber in the aforementioned EP-A-0 328 193. In each of the drawing figures there is illustrated a differentially controllable valving arrangement for controlling the thrust on the piston 13 including adjustable set screw 109 having a conical end surface 111 variably spaced from a similarly shaped seat 113 for supplying air from the pressurized source to the air control valves to compensate for variations in external forces opposing piston motion. Set screw 109 may be adjusted to vary the restriction between chamber 33 and channel 115 leading to control valve 15. The corresponding channel 117 leading to control valve 17 has a fixed restriction. The restriction tends to be self adjusting in the sense that if piston motion is opposed then the pressure driving the piston increases tending to correct for the increased opposition. Figures 10 and 11 are similar to Figure 1, but each illustrates a scheme wherein the pneumatic damping means is differentially adjustable to vary piston deceleration as the piston approaches one extremity relative to piston deceleration as the piston approaches the other extremity. The pneumatic damping means includes a volume varying adjustable member in Figure 10, and, in Figure 11, an adjustable member for controlling air wscape from the pneumatic damping means. In Figure 10, a pair of adjustable set screws 119 and 121 seal corresponding holes leading to the chambers 36 and 35 respectively. Axial movement of one of these screws varies the volume of the piston motion damping chamber. When the piston is near the end of its travel, this small volume becomes a significant part of the total volume of the damping chamber and a change in that volume has a significant effect on the chamber pressure and, therefore, on the damping force. For example, if set screw 121 is withdrawn increasing the volume of chamber 35, the opening of reed valve 101 (at peak or source pressure) will be delayed until the piston is closer to its rightmost position. A fine tuning of the damping motion at one extreme of piston travel relative to damping at the other extreme is therefore possible. Such a fine tuning may also be achieved by bleeding air from the damping chamber as in Figure 11 rather than varying the volume of that chamber as in Figure 10. In Figure 11, a pair of needle valves 123 and 125 control air seepage from the damping chambers, thereby controlling the time at which peak pressure occurs. Little has been said about the internal combustion engine environment in which this invention finds great utility. That environment may be much the same as disclosed in the abovementioned copending applications and the literature cited therein to which reference may be had for details of features such as electronic controls and air pressure sources. In this preferred environment, the mass of the actuating piston and its associated coupled engine valve is greatly reduced as compared to the prior devices. While the engine valve and piston move about 11.4 mm (0.45 inches) between fully open and fully closed positions, the control valves move only about 4.44 mm (0.175 inches), therefor requiring less energy to operate. The air passageways in the present invention are generally large annular openings with little or no associated throttling losses.

Claims: A pneumatically powered valve actuator comprising a valve actuator housing (19), a piston (13) reciprocable within the housing (19) along an axis, the piston having a pair of oppositely facing primary working surfaces (38), a pressurized air source (33; 39), a pair of air control valves (15; 17) movable between open and closed positions, means (25; 27) for selectively opening one of said air control valves (15; 17) to supply pressurized air from the air source (33; 39) to one of said primary working surfaces (38) causing the piston (13) to move and pneumatic means (35; 37; 43; 67) for decelerating the piston (13) near the extremities of its reciprocation, characterized in that the pneumatic means (35; 37; 43; 67) includes a one-way pressure relief valving arrangement (101; 103) for venting air from the pneumatic means (35; 37; 43; 67) to the pressurized air source (33; 39). The pneumatically powered valve actuator of Claim 1, wherein the pneumatic means (35; 37; 43; 67) is differentially adjustable to vary piston deceleration as the piston (13) approaches one extremity relative to piston deceleration as the piston (13) approaches the other extremity. The pneumatically powered valve actuator of Claim 2, wherein the pneumatic means (35; 37; 43; 67) includes a volume varying adjustable member (119; 121). The pneumatically powered valve actuator of Claim 2, wherein the pneumatic means (35, 37; 43; 67) includes an adjustable member (123; 125) for controlling air escape from the pneumatic means (35; 37; 43; 67). The pneumatically powered valve actuator as claimed in one of the Claims 1 to 4, wherein the one-way pressure relief valving arrangement comprises a plurality of reed valves (101; 103). The pneumatically powered valve actuator as claimed in one of the preceding Claims, characterized by a differentially controllable valving arrangement (109, 115, 117) for supplying air from the pressurized air source (33; 39) to the piston (13) to compensate for variations in external forces opposing piston motion.

Inventors: ERICKSON FREDERICK LOGAN, RICHESON WILLIAM EDMOND, ERICKSON, FREDERICK LOGAN, RICHESON, WILLIAM EDMOND
Applicants: MAGNAVOX ELECTRONIC SYSTEMS, MAGNAVOX ELECTRONIC SYSTEMS COMPANY
Application year: 1989
Main classification: F01L9
IPC codes: F16K31, F03G7, F01L9