US2962013A - Engine valve mechanism - Google Patents

Description

Nov. 29, 1960 F. c. REGGIO 2,962,013

ENGINE VALVE MECHANISM Original Filed Jan. 20. 1943 5 Sheets-Sheet 2 v Nov. 29, 1960 F. c. REGGIO ENGINE VALVE MECHANISM Original Filed Jan. 20, 1943 5 Sheets-Sheet 3 Nov. 29, 1960 F. c. REGGIO 2,962,013

` ENGINE VALVE MECHANISM Original Filed Jan. 20, 1943 5 Sheets-Sheet 5 IN VENTOR United States Patent Q F ENGINE VALVE MECHANISM Ferdinando C. Reggio, Tampa, Fla. (P.0. Box 692, Norwalk, Conn.)

Application Nov. 8, 1952, Ser. No. 319,577, which is a division of application Ser. No. 472,947, Jan. 20, 1943, now Patent No. 2,621,640, dated Dec. 16. 1952. Divided and this application Dec. 13, 1957, Ser. No. 702,607

11 Claims. (Cl. 123-92) This invention relates to lluid pressure actuation of reciprocating members, and more particularly to mechanisms for hydraulically operating the valves of internal combustion engines and the like, and the plungers of engine fuel injection pumps. The instant application is a division of application Serial No. 319,577 tiled November 8, 1952, now Patent No. 2,816,533 which in turn is a division of application Serial No. 472,947, filed January 20, 1943, now Patent No. 2,621,640, issued December 16, 1952.

One object of the invention is to provide a light, compact and simple valve mechanism.

Another object is to provide a hydraulic mechanism permitting a considerable increase in the valve velocity.

A further object resides in the provision of a valve mechanism utilizing lubricating oil or other suitable hydraulic med-ium for actuating and cooling the engine valves.

Still another object is to provide a hydraulic valve mechanism which permits a considerable reduction in the weight and size of the engine.

A further object is to provide, in substitution for the valve springs, means for applying a fluid pressure load to the valves tending to close the same.

A still further object is to provide means for automatically altering the valve timing with changes of engine operative conditions.

Another object resides in the provision of a valve mechanism wherein the number and weight of the reciprocating parts is considerably reduced.

A further object resides in the hydraulic actuation of the plunger of a fuel injection pump.

Other objects and advantages of the invention will become apparent from the following description taken in connection with the accompanying drawings, in which:

Figure l is a fragmentary vertical section showing one of the preferred embodiments of the invention applied to a radial air cooled aircraft engine.

Figure 2 is a fragmentary section through the valve of Figure l.

Figure 3 is a fragmentary front view of the engine of Figure l.

Figures 4 and 5 are `diagrams indicating the valve timing under different engine operative conditions.

Figure 6 is a section illustrating a second preferred embodiment of the invention applied to a liquid cooled engine.

IFigures 7 and 8 are sections along the lines 7-7 and 8-8 of Figure 6, respectively.

Figure 9 shows a detail of Figure 6 in larger scale.

Figure 10 is a section illustrating =a third preferred 2,962,013 Patented Nov.. 29 1960 ice embodiment of valve and fuel pump actuating mechanism according to the invention, applied to a liquid cooled engine.

Fig. 11 is a section of a detail of Figure 10 in larger scale.

Fig. 12 is a diagrammatic view of the mechanism of Figure 10 when applied to a six cylinder in-line engine.

Fig. 13 is a fragmentary section through a rotary control valve for a nine cylinder radial engine.

Fig. 14 is a section along line A-A of Figure l.

Figures l5 and 16 indicate diagrammatically a conventional arrangement of torque responsive device or hydraulic torque meter. Y

The illustrative example of embodiment shown in Fig-f ure 1 includes conventional parts of arad'ial air cooled aircraft engine which will easily be recognized by those skilled in the art, namely a portion of cylinder head 13, a valve 14, and a cam 15 mounted within the engine housing 16 for actuating the valves. The valve 14 is slidably mounted in a guide 21 above which there is piro-x vided a cylinder 20 having a bore coaxial with the guide 21 and cooling fins 19. The cylinder 20, preferably carried by or attached to the guide 21, may be secured to the cylinder head 13 by screws, not shown. A piston 18, closely fitting the cylinder 20 and slidable therein, is attached to the stem of the valve 14 by Vmeans of a nutV and lock nut. The annular chamber 23 below the piston 18 is connected by means of portV 24 and pipes 25, 26 with an annular conduit 28, while the cylinder chamber 31 above piston 18, closed Aby a cover 22, is connected by means of port 32 and pipe 34 with the upper chamber 44 of a differential cylinder 36 having slidable therein a differential plunger 38 actuated through a roller 40 from the cam 15. The plunger has an upper portion of smaller diameter slidable in the bore 41 of the cylinder 36,`and a coaxial lower portion of larger diameter slidable in the bore 42, thus providing an annular chamber 45 connected by means of port 46 and duct 48 with the annular`| conduit 28.

The plunger 38 is provided with an axial bore 50 where-v in there is slidably mounted a piston 51 upwardly urged by a calibrated spring 52. A ring 53 seated in la groove formed in the bore 50 limits the upward displacement of piston 51. The space below the latter communicates by way of a drain conduit 55 with the engine crankca'se. An annular groove 56 formed in the plunger 33 between chambers 44 and 45 is connected by way of duct 58 with the conduitSS. A small orice 54 may be provided in the piston 51sto establish flow communication between chamber 44 and drain conduit 55. An axial groove 60 formed in the lower cylindrical portion of the plunger 38 communicates by way of a duct 61 formed in the latter with the chamber 44 through an orcecontrolled by a ball or check valve 62 permitting oil How from the groove 60 to the chamberA 44 but preventing the return ow from saidtchamber back to the groove 60. The latter also communicates with an annular duct 64 which receives lubricating oil under pressure from an engine pump 65 through a pipe 66, a flow-restricting calibrated orifice 68 and a pipe 70. Connected with the latter there is provided a pressure regulating plunger valve 72 controlling a discharge orifice 74 and loaded by a calibrated spring 7 3 mounted between said valve and a spring seat carried by a slidable rod 75 coaxial with the spring 73, whereby the load of the latter and in turn the oil pressure in the duct 64 are determined by the axial adjustment of said rod 75.

The rod 75 extends through the wall of a housing 76 and is attached to the movable wall of a bellows 78 contained therein and connected with a spring 80 which may be adjusted by the operator by means of a manual control 82 and a suitable linkage 83. A pipe 84 connects the engine induction manifold 85 with the space within the bellows 7S; and the space comprised between the housing 76 and the bellows 7,8 may be connected by means of a conduit 86 and a three-way cock 88 either with the pipe 90 open to the atmosphere, or with pipe 91 connected with the engine exhaust manifold 92. With the cock 88 as shown in Figure 1 the bellows 78 is surrounded by the exhaust pressure; and if the cock is rotated 90 degrees clockwise the bellows 78 will be sur rounded by atmospheric pressure.

An engine driven pump, preferably of the high-pressure type having two or more stages and diagrammatically represented at 93, receives oil from a suitable reservoir, sump, or other source, for example from the discharge line 66 of the engine-driven lubricating oil pumpl 65, and through pipe 94 supplies high pressure oil to the annular conduit 28 and to the oil accumulator and pressure regulator cylinder 95 having slidable therein a piston 96 loaded by a spring 98. The piston controls axially spaced orifices 100 and 101 through which the excess oil is discharged into a low-pressure return conduit 102 which in the example shown in Figure l is connected with the intake side of pump 93. A hand pump 103 may be provided to supply pressure oil to the accumulator 95, and thence to the annular duct 64, through pipe 104, and prime the hydraulic system before starting the engine.

The oil discharge of the volumetric engine-driven pump 93 is substantially proportional to the engine speed. When the engine is operating within a predetermined speed range the larger orifice 101 is covered by the piston 96 and said oil discharge, minus the various leakages of oil which are negligible as compared with the output of the pump 93, is forced through the small calibrated orifice 100. The oil velocity through this orifice is thus substantially proportional to the engine speed, and the corresponding drop of pressure across the orifice 100 is therefore proportional to the square of the engine speed, thus maintaining the oil pressure in the annular conduit 28 at a value which is also substantially proportional to the square of the engine speed. If such speed decreases below a predetermined lower value, the piston 96 covers part of the orifice 100 and prevents the oil pressure from dropping below a corresponding minimum pressure limit. Conversely, if the engine speed increases beyond the maximum rated speed the piston 96 uncovers the orifice 101 and prevents the oil pressure in the conduit 28 from increasing above a preselected corresponding maximum pressure limit.

In the preferred embodiment of the invention, the orices 100 and 101 are so spaced and the spring 98 is so designed as to obtain the following operation. When the engine is stationary, the` piston 96 is in its extreme left position and covers both orifices 100 and 101. As soon as the engine is cranked, the pump 93 discharges oil into the various lines and in the cylinder 95. Since no oil outlet is provided, the oil pressure builds up and moves the piston 96 to the right until the orifice lili) is partly uncovered. As the capacity of the pump 93 is considerably larger than the various oil leakages on the discharge side of the pump even at very low cranking speed, as soon as the engine crankshaft is turned, the piston 96 assumes such a position as to partially open the discharge orifice 100. Under these conditions, the specific pressure of the oil on the discharge side of the pump 93 will of course be equal to the load of spring 98 (corresponding to said position of piston 96) divided by the piston area.` It follows that even at the lowest cranking speed of the engine, enough oil pressure is maintained to properly operate the hydraulic valve gear, as set forth hereinafter.

As the cranking speed of the engine increases, the piston 96 moves further to the right to increase the open area of orifice 100, until the latter is completely uncovered. In the preferred embodiment of the invention this happens just when the engine operates at idling speed. It is therefore apparent that for all values of engine speed from a few r.p`;m.s up to idling speed the position of the piston 96 varies merely by an amount equal to the diameter of the small orifice 100. As a result, the load of spring 98, and the oil pressure which is proportional to the spring load, are approximately constant for all engine speeds up to idling speed. Within such a speed range the spring-loaded piston 96 and the orifice 100 operate like a conventional pressure regulating valve, uncovering more or less of the area of orifice 100` to accommodate the variations in the delivery of the volumetric pump 93, which delivery is proportional to engine speed, while the oil pressure upstream of the orifice and the velocity of flow of the oil through the uncovered portion of the orifice 100 remain approximately constant.

As stated above, at idling speed the position of piston 96 is such as to fully uncover the orifice 100. On the other hand, the larger orifice 101 is so located as to be uncovered by the piston 96 only when the engine speed exceeds the maximum designed engine speed. Thus for all values of engine speed comprised between idling and maximum speed, that is to say, throughout the normal operating range of engine speeds, the orifice 100 is fully open, while the larger orifice 101 is completely covered by the piston 96; and all of the oil delivered by the volumetric engine-driven pump 93 is discharged through the fully uncovered orifice 190. Thus for all engine speeds within said operating range the velocity of oil flow through orifice 100 is proportional to the engine speed, and it follows of course that the pressure differential across the orifice 100 (or relative pressure in the cylinder is proportional to the square of the engine speed.

The annular chamber 23 in the valve actuating cylinder 20, connected with the conduit 28, serves to apply to the piston 18 an upward oil pressure load tending to close the valve 14. Within the preselected engine speed range said load is proportional to the square of the engine speed and therefore proportional to the inertia of the valve. The conventional spring or springs connected with the engine valve may thus be eliminated. In similar manner the annular chamber 45, also connected with the conduit 28, exerts on the plunger 38 a downward oil pressure load which acts against the inertia of said plunger to maintain at all times the roller 40 thereof in contact with the cam 15.

As the engine operates, oil is displaced at each cycle to and from the small annular chambers 23 and 45, andA oil pressure pulsations occur which keep the piston 96 of the pressure accumulator 95 in continuous motion. In an engine having a large number of valves at different phases these pressure surges are very considerably reduced. In order further to weaken the intensity of these surges, pressure cushioning means such as chambers containing a compressible uid or chambers having yielding wall means may be provided in connection with any suitable part of the hydraulic system, for example near the valve cylinders 20 or the differential plungers 38, or the conduit 2S, as Will be obvious to those skilled in the art. An example of such a cushioning means is illustrated in Figure l in the form of a cylinder 111 having a chamber 114 therein connected with the conduit 28 and defined in part by a slidable piston 112 loaded by a spring 113. The space at the opposite side of the piston 112 is connected by means of a duct with the drain pipe 110.

Lubricating oil leaking from the high-pressure chamber 23 between the valve stem and the guide 21 is trapped in the groove 106 and led back to the engine sump by Way of the drain pipes lilS and 110. Oil upwardly leaking from the chamber 45 is trapped in the groove 56 and returned to the sump by means of the ducts 58 and 55.

Figure 1 shows the valve 14 in closed position, the plunger 38 in its lowermost adjustment, and the small piston 51 in intermediate adjustment between its uppermost and lowermost positions relative to the plunger 38, the load of the piston spring 52 being balanced by the oil pressure in chamber 44. As the plunger 38 is lifted by the rotating cam the oil pressure in said chamber increases and displaces the piston 51 downwardly against the load of spring 52, whereupon oil is forced through pipe 34 into the chamber 31 and lifts the valve 14. The deceleration and return stroke of the latter are determined by the upward load exerted on the valve piston 18 by the high pressure oil contained in the annular chamber 23. When the valve 14 closes the plunger 38 is still in its return stroke, the oil pressure in chamber 44 drops to a low value, and the spring 52 expands and lifts the piston 51. However during this cycle some oil has escaped through the small orifice 54 and various other oil leakages have occurred, and the piston 51 will therefore attain a higher position relative to the plunger 38 than it had at the beginning of the cycle. The oil pressure in chamber 44 will accordingly be lower than in the duct 64, causing the check valve 62 to open and admit a certain amount of oil into chamber 44.

It is thus apparent that the displacement of the piston 51 has the effect of rendering a portion of the stroke of plunger 38 inoperative, causing the valve 14 to open later and close earlier. If the pressure in the annular duct 64 is sufficiently low the piston 51 will attain its uppermost position and rest against the ring 53 when the plunger 38 is at rest between lifts, and the valve 14 will remain open during the minimum designed duration. Conversely, if the pressure in conduit 64 is maintained at suiciently high Value so as to keep the piston 51 constantly in its lowermost position, the latter becomes ineffective and the period during which the valve 14 remains open reaches its maximum designed value. By this arrangement the valve timing may be controlled While the engine is in operation as shown diagramm-atically in Figures 4 and 5, wherein b and c indicate the intake and exhaust phases respectively, and a represents the scavenging phase during which both intake and exhaust valves are open. The former figure, corresponding to the higher value of the oil pressure in the annular duct 64, obtained when the rod 75 is in its extreme left adjustment, shows a considerable scavenging phase `or valve overlap, during which the cylinder may be swept by the air or mixture from the supercharger. Conversely, if the rod 75 is adjusted in its extreme right position, the oil pressure in duct 64 is kept at the lower designed limit, and the scavenging phase is reduced to zero or to a negligible value as shown in Figure 5.

The bellows 78 cooperating with the spring 80 actuates the rod 75 to increase or decrease the duration of the scavenging phase or valve overlap as the pressure in the engine induction manifold 85 becomes higher or lower than that in the exhaust pipe 92. With the three-way cock 88 rotated 90 degrees clockwise, the bellows 78 becomes responsive to the difference between engine manifold pressure and surrounding atmospheric pressure, and operates the rod 75 to reduce the valve overlap to a negligible value when said pressure difference becomes negative. The return flow through the engine cylinder at the end of the exhaust stroke from the exhaust pipe 92 to the induction manifold 85 when the pressure in the latter is low, as when the engine is idling, may thus be avoided. The valve timing may also be adjusted by the operator .While in flight by means of the control member 82.

In Figure 2 there is indicated a valve 14 having a hollow stem or cavity 115. A rod 116 carried by the cover 22 extends therein leaving an annular interspace through which during engine operation oil ows back and forth at high velocity and thus effectively cools the valve by transferring heat by conduction and convection from the valve head to the stern portion thereof, in substantially the same way as in the sodium-cooled valves which are extensively used in aircraft engines. A helical groove 117 is provided in the rod 116, and ports 24 and 32 are preferably directed tangentially with respect to the cylinder 20 as shown in Figure 14 to impart rotary motion to the oil therein and further improve the cooling of the valve 114. Oil cooling devices, such as fins 19 and 30, are provided in connection with the cylinder 20 and the oil conduits 24 and 25.

Figure 3 diagrammatically illustrates the various hydraulic connections in the valve mechanism of a radial aircraft engine, wherein the cylindrical chamber 31 of each valve cylinder 20v is connected with the lcorrespond,- ing chamber 44 of a plunger housing 36 by means of a pipe 34, while the annular chambers 23 and 45 are all connected by means of pipes 25-26 and 48, respectively, with the annular conduit 28 which is provided with one or more pressure cushion chambers 111 and receives high pressure oil through the line 94.

The mass of valve 14 with piston 18 attached thereto is only a fraction of the total mass of a conventional valve with the reciprocating mechanism connected therewith, and accordingly the load exerted on the piston 18 by the oil pressure in the annular chamber 23 need be only a small fraction of the load of the conventional valve springs. If high pressure oil is employed, the Vbore of cylinder 20 may be made only slightly larger than the stern of the valve 14. Furthermore, owing to the fact that only axial load is applied to the valve and that the latter is cooled by means of lubricating oil, the valve guide 21 may be made quite short. The cylinders 20 may therefore be made very small as compared with the relatively bulky conventional rocker housing, and the frontal area and the weight of the engine may be considerably decreased. In addition, the cylinder head becomes much simpler, and more space in the immediate vicinity of the combustion chamber becomes available for cooling fins.

In the above described mechanism each valve is actuated from a corresponding cam-actuated plunger by means of a substantially incompressible liquid column, the valve lift being determined by the cam profile. Figure 6 illustrates another preferred embodiment of the invention wherein an engine-driven rotary control valve is substituted for said cam and plunger. The cylinder head 121 of a liquid-cooled, valve-in-head engine is provided with intake and exhaust valves 122 and 123. The latter valve has attached thereto a piston 124 slidable in a cylinder 125 attached to and coaxial with the guide 126 wherein there is formed a groove 128 connected with a drain duct 130. The upper end of the cylinder 125 is closed by a threaded cover 131 and the chambers 140 and 132 on opposite sides of the piston 124 communicate by means of ports 142 and 133, shown in Figure 9, with the pipes 143 and 134 respectively. Similarly, the intake valve 122 carries a piston 124 slidable in a cylinder having upper and lower chambers and 132 communicating with pipes 167 and 134' respectively. Said pipes 134 and 134 are connected with a liquid cooled conduit 135 communicating by means of a duct 136 with the discharge side of a high pressure engine-driven oil pump 137 diagrammatically indicated as a two-stage gear pump, and with an oil accumulator and pressure regulator 138 similar to the unit 95 and having similar functions, namely to return the excess oil to the inlet side 139 of the pump 137, to maintain the oil pressure substantially proportional to the square of the engine speed between predetermined maximum and minimum, values thereof, and to cushion the pressure surges in the high-pressure system. The oil pressure in chambers 132 and 132' constantly exerts on -the valves V123 `and 122 a load tending to close the latter.

'The pipes -167 `and 143 are connected with ports 166 and 165 opening in the upper and lower parts of a cylindrica'lbore 145 formed in the housing 144 of a rotary control `valve having a rotor 155 therein slidably mounted on a splined portion 152 of a shaft 146 journaled in the lower valve cover 1-48 and carrying a pinion 150 meshing with an engine-driven ygear 151. The upper portion 153 of the shaft 146 is journaled in and extends through the upper cover 154 of the valve. In this valve arrange ment, designed for a four cycle in-line engine, the shaft -146 revolves at half crankshaft speed. The rotor 155 is provided with a groove 156 engaging a lever 157 connected with an external lever 160 whereby by turning the latter the axial adjustment of the rotor 155 .relative to the ports 166 and 165 may be varied. A groove formed in the central portion of the rotor provides an annular chamber 162 which by means of a conduit 163 is kept in `constant communication with the discharge side of the pump 137. The upper and lower portions of the bore -145 are connected with a drain conduit 164 for leading low pressureioil `back to the engine sump or other suitable reservoir.

The rotor 155 in its upper portion is provided with two grooves 168 and 170, the former communicating with the groove 162 and the latter with the low pressure drain conduit 164. As the rotor S revolves these two grooves 168 and 17.0 successively register with the port 166 thereby successively connecting the chamber 140' with the high pressure side of the pump 137 and with the drain conduit. In similar manner the lower portion of the rotor 155 is provided with two grooves 171 and 172 adapted successively to register with port 165, the former groove communicating with the groove 162 and Vthe latter with the drain conduit 164, so that in operation the cylinder chamber 140 is also successively connected with the high and low pressure oil lines 163 and 164. The controlling edges dening said four grooves 168 and 170 to 172 are, in part at least, not parallel to the axis of the rotor but are formed in such manner as to vary the duration of the intervals during which the ports 165 and 166 are connected with the high and low pressure oil lines when the axial adjustment of the rotor 155 is changed.

With the rotor '155 in the angular position shown in Figures 6 to 8 the cylinder chambers 146 and 140 are Yconnected by means of lines 143 and 167, ports 165 and 166, and grooves 170 and 172, respectively, with the low pressure drain conduit 164, and the valves 123 and 122 `are kept closed by means of the high pressure oil contained in the annular chambers 132 and 132l respectively. As the rotor 155 revolves in the direction indicated by the arrows, the connection between port 165 and groove 172 is interrupted and connection between said port and the groove 171 is established, and high pressure oil from the pump 137 and the accumulator 13S is admitted through the pipe 143 to the cylinder chamber 140, thus determining downward acceleration of the diiierential valve piston 124. As the resultant hydraulic load applied to said pis- `ton `is approximately constant during the greater part of the lift, the valve acceleration will accordingly be approximately constant.

As shown in Figure 9, at the beginning of the valve lift `only the small portion of port 142 above the dotted line B is open, the remaining portion thereof being covered by `the piston 124. However the valve velocity is low at the very beginning of the lift, and even the small eiective area of the port is sufficient for admitting the necessary volume `of oil without causing any appreciable drop of pressure Yinchamber1140. As the lift and the velocity of the valve 123 increase, the piston 124 uncovers the port 142 and the effective area thereof attains its maximum designed walue. At .the same time oil is displaced from `the .an-

nular chamber 132 through the port 133; `and as .toward the end of the :downward stroke of the valve the piston 124 approaches and attains its lowermost position indicated in dotted line at 124 and covers theportion of said portabove the dotted dine C, the effective open area `of the latter decreases, determining a rapid increase of oil pressure in .chamber 132 and causing quick deceleration of the valve so as Ato reduce the impact speed of the piston 124 against the end wall of the cylinder 125 to a suitable value. The high pressure in chamber keeps the valve in open position until 4the communication between the port '165 and the high pressure groove 171 is interrupted and the connection between said port and the low pressure groove 172 is established; whereupon the pressure in chamber 140 drops and the high pressure constantly existing in the annular chamber 132 causes upward displacement ot the valve 123. Toward the end of this stroke, owing to the decrease in the effective area of port 142, the oil pressure .in chamber 140 rises and causes rapid deceleration of the valve 123 so as to reduce the speed of impact of the latter against the valve seat to an appropriate value. Thereafter the valve is kept closed by the oil lpressure-in chamber 132 until high pressure oil is again admittedzto the chamber 140 by the revolving rotor 155.

The intake `valve `122 is actuated by means of the upper part of the rotor 15S, controlling the `port 166, in similar manner, and it is therefore regarded as unnecessary again to describe `its operation in detail.

The oil .delivery of the engine-driven pump 137 is proportional to the engine speed and exceeds the total oil displacement of the valve cylinders 125 which is also proportional to the engine speed. The excess oil, which thereforeis also proportional to the engine speed, returns to the inlet `side 139 of the pump through an orice 176 which may initially be adjusted by means of a needle valve 177. It follows that the oil velocity through the orifice V176 is proportional to the engine speed, and the corresponding pressure differential between upstream and downstream sides of orifice 176 is proportional to the square of .the engine speed; that is to say, the oil in the high pressure-system, including chambers 132 andconduits 135, 136 and 163, is maintained at a pressure proportional to the square ofthe engine speed whenever the engine is operating within a predetermined speed range. The upper ylimit of said range is attained when, under abnormally highengine speed, the resiliently loaded plunger which is slidably mounted 'in cylinder 138 uncovers a large port 178 Iand thereby prevents further increase of oil pressure. ihe lower limit of said range is attained at a predetermined low engine speed, for example engine idling speed, las the plunger covers the orice 176 in part so as to reduce the effective area thereof and thereby prevents further drop of oil pressure. The actuating hydraulic load applied to the Valves by means of this pressure and the positive acceleration thereof are therefore proportional to the square of the engine speed, and the valve velocity is directly proportional to the engine speed. Thus the curve representing the valve lift versus the corresponding angular position of the crankshaft is independent of the engine speed, exactly as if the valves were reciprocated from a cam. In the present case, however, the valve velocity may be much higher than in the conventional valve mechanism, as the valve deceleration and inertia are not limited by the load of the valve spring. Furthermore the mass of the valve and piston 124, only reciprocating parts, and the mass of the oil column movable therewith are only a fraction of the lmass of the conventional valve -with the operating gear connected therewith.

With the rotor having high and low pressure grooves 168 and 170 to 172 as illustrated in Figure 6 the valve timing and the valve overlap or scavenging phase `may be varied by altering the angular position of the lever and thereby the axial adjustment of the rotor 155 relative .to .the .ports and 166. A speed responsive accepts 9. device including centrifugal'flyballs 180 driven from the engine by means of gears 151 and 181 actuates a resiliently loaded sleeve 182 connected with a floating lever 183 and displaces the right end of the latter downward or upward upon an increase or a decrease of the engine speed, respectively. An intermediate point of lever 183 is pivoted to a rod 185 secured to a piston 186 loaded by a spring 188 and slidable in a cylinder 190 connected by means of an oil conduit 191 with a conventional hydraulic torque meter diagrammatically indicated in Figures 15 and 16 whereby the piston 186 is displaced upward or downward upon an increase or a decrease of engine torque, respectively. The left end of lever 183 is connected by means of a rod 184 with the lever 160 so as to vary the engine valve timing in dependence upon changes of engine speed and torque, and in particular t reduce the valve overlap as the engine speed, or torque, or both decrease. Torque meters as indicated in Figures l and l6 are well-known in the art. They often include a hydraulic torque responsive device connected with a planetary speed reduction gear arranged between the engine crankshaft and the propeller shaft, for maintaining the oil pressure in suitable conduit and reservoir means or pressure line 191 at all times proportional to the torque transmitted from the engine to the propeller. Figure l5 shows a section through the nose of the engine of Figure 1 perpendicular to the shaft thereof, wherein 315 indicates the engine nose housing containing a reduction gear of the planetary type having planet pinions 316 carried by journals 317 supported by an annular member, not shown, rotatable with the propeller shaft 318 and engaged between a sun gear 319 secured to the engine crankshaft and an outer ring gear 320. The latter is connected by way of rod 321 with a piston 322. An engine driven pump 323 discharges oil under pressure into chamber 324 0n the outer side of the piston 322 for applying to said piston a load balancing the tangential load, proportional to the propeller torque, transmitted by the outer ring gear 320 to the rod 321. Oil escapes from the chamber 324 to the low pressure chamber 325 by way of a duct 326 and a restricted passage of variable effective area 327 between the cover 387 and the piston 322, whereby the effective area decreases or increases upon a displacement of the piston 322 to the left or right, respectively. The torque transmitted through the reduction gear tends to rotate the outer ring gear 320 counter-clockwise and displace the piston 322 toward the left, and is normally balanced by the oil pressure in the chamber 324. An increase of torque causes displacement of the piston 322 toward the left thereby reducing the eective area of the restricted passage 327 and increasing the oil pressure in the chamber 324 in proportion to the increase of torque, whereupon the equilibrium of the piston 322 is reestablished. The cylinder 190 of the regulating device, as shown for instance in Figure 6, includes a pressure responsive element such as a piston 186 connected by means of the pressure line 191 with the pressure chamber 324 of the torque meter so that the pressure below the piston 186 is always proportional to the engine torque, and therefore the adjustment of said piston and the valve timing are caused to vary as a function of said torque.

If desired, the oil pressure in the high pressure system may be maintained constant regardless of the engine speed by closing the orifice 176 by means of the needle valve 177.

The valve mechanism according to the invention may be used for actuating reciprocating valves of compressors, pumps and the like, as will be obvious to those skilled in the art, and also for actuating other reciprocable members such for example as the plungers of fuel injection pumps as illustrated in Figure l0, wherein the cylinder head 200 of an engine is provided with a passage 202 in which there is mounted a fuel injector 201 having a nozzle part 203 clamped on a seat 204 and adapted to intermittently discharge fuel under pressure into the cylinder. Mounted in coaxial bores formed in the injector housing 201 there are a rotatable pinion 205 provided with external gear teeth meshing with `a fuel control rack 224 transversely slidable with respect to the axis of said pinion, and a sleeve 206 having at least one port 208 connected with a groove 210 to which fuel is led by a duct 211 from a port 212 adapted for connection with a fuel line 212. sleeve 206 are tightly clamped by means of a threaded annular retaining member 214. A plunger 215, slidaable in the cylindrical bore formed in the sleeve 206, pinion 205 and injector housing 201, has an inclined or helical scroll 216 for controlling the port 208 so as to vary the fuel discharge of the injector upon changes of angular adjustment of said plunger. A piston 218 of larger diameter than plunger 215, formed integral with or secured to the upper end of the latter, slidably ts in a cylinder machined in the upper portion of the housing 201 and closed by a threaded cover 221, thus providing a chamber 228 and an annular chamber 230 above and below the piston 218, respectively.

The plunger 215 is shown in its uppermost position, with the piston 218 in contact with the cover 221, in which the fuel pressure chamber 220 is in communication with the annular fuel groove `210. As the plunger 215 is forced downward, it closes the port 208 and forces fuel from the chamber 220 through the nozzle 203 into the engine cylinder. plunger 215 is attained when the piston 218 reaches the lower end of the chamber 230. A longitudinal slot 222 formed in the intermediate portion ofthe plunger 215 is engaged by the inner end of a pin 223 forced into a radial bore formed in pinion 205 through a suitable opening provided in the housing 201, not shown in the drawing, whereby the angular adjustment of the plunger and in turn the fuel discharge of the injector are determined by the axial adjustment of the control rack 224, as is well known in the art. The injector 201 is secured to the engine by means of fastening devices such as screws 225.

The chambers 228 and 230 are connected by means of ports 233 and 237, which may be similar to the ports 142 and 133 of Figure 9, with the pipes 231 and 232 respectively. The latter pipe is connected with the high pressure oil conduit 135', similar to thev conduit 135 of Figure 6 whereby an upwardly directed pressure load is. constantly exerted on the piston 218. The former pipe 231 is connected -with a port 234 opening in the bore 235 of a rotary control valve 236 including a rotor 238 slidably mounted on a splined shaft 240 journaled in the lower cover 241 and driven from the upper end 153 of the shaft 146 (shown in Figure 6) of the control valve 144 by means of a coupling 242. The rotor 238 has a groove 243 constantly communicating with the high pressure oil conduit 163', similar to conduit 163 of Figure 6. A cover 244 closes the bore 235 at the upper end thereof, and a low pressure drain conduit 245 is connected with the upper and lower ends of bore 235. The rotor 238 has two grooves 246 and 248 communicating with the high and low pressure lines 163 and 245 respectively, and as the rotor 238 revolves said two grooves successively register with the. port 234 and cause the plunger 215 to reciprocate in substantially the same manner as described in detail in connection with the valves 123 and 122. As the controlling edges of the grooves 246 and 248 are not parallel to the axis of the rotor 238, the injection timing may be varied by altering the axial adjustment of said rotor relative to the port 234. To that end the rotor 238 is provided with a groove engaged by a lever 253 connected with an external lever 254 actuated by way of the rod 255 either manually or by means of engine operative condition responsive means such as the speed responsive device 180.

The valves 122 and 123 in Figure l0 are the same as in Figure 6, while the valve actuating cylinder 260' or 260, the latter of which is shown in section in Figure 11,'

The nozzle part 203 and they The lowermost position of the.

has an upper port 261 including a main circular portion and a comparatively very small triangular portion so de. signed that when the upper edge of the valve piston 124 approaches its uppermost position indicated by the dotted line D, the effective open area of the orifice 261 becomes suiciently small to cause quick deceleration of the valve before the latter reaches its seat, thus rendering possible the adoption of valve velocities many times as great as those employed in conventional high-speed engines. However with such form of port the initial part of the valve lift may be slowed down by the throttling effect of the small initial effective area of the port To obtain a high value of deceleration prior to the seating of the valve as well as unrestrained initial valve lift acceleration, a second communication is provided between the pipe 143 and the upper part of the cylinder chamber 140 by means of conduits 262, 263 and 265, including a check valve 264 permitting flow of oil from the pipe 143 into the upper chamber but preventing flow of oil in the opposite direction. A similar arrangement including a check valve 266 is provided at the lower end of the annular chamber in connection with the lower part 268,V whereby a high value of deceleration may be obtained at the end of the valve lift as well as unrestrained acceleration at the beginning of the valve closing stroke.

The sections of the rotary control valve 144 in Figures 7 and 8 indicate six equally spaced radial ports in a plane perpendicular to the axis of the rotor 155. The latter rotates in the direction of the arrows at half crankshaft speed. This arrangement is intended for a six cylinder four-cycle engine. Figure 12 diagrammatically indicates the hydraulic connections, by means of pipes 167 and 143, between the rotary control valve 144 (the upper fragment of which is shown in Figure l) and the intake and eX- haust valves 260' and 2 60, respectively, as Well as the connections, by means of pipes 231, between the rotary control valve 236 and the fuel injectors 201 of a six-cylinder engine 265 wherein the ignition order is 1-4--2- 6-3-5. It will be noted that each cylinder is provided with two exhaust valves 260 connected with the rotary valve 144 by means of a common pipe 143, and with two intake valves 260 connected with said rotary valve through a single line 167.

It will be obvious to those skilled in the art that a valve mechanism as shown in Figure l, wherein the valves are operated by means of cam-actuated plungers, mav be used in connection with an in-line enginge as wellA as with a radial engine; and conversely theA valve mechanisms illustrated in Figures 6 and 9 to l1, employing a rotary control valve, may be applied to a radial engine justas well as to an in-line engine. Thus Figure 13 indicates a transversal section through the upper portion of a rotary control valve 270 for a nine cylinder air-cooled radial aircraft engine, having nine equally spaced radial ports 2.71, 272, 273, etc., connected with the intake valves of the engine cylinders. Said ports are controlled by a rotor 275 driven by a splined shaft 276 at one-eighth crank,- shaft speed in counter-clockwise direction, The, rotor 275 is provided with four equally spaced grooves 27'8 connected with the high pressure oil system, alternating with four grooves 280 communicating with the oil drainV line. The rotary valve is otherwisesubstantially similar to valve i144 and it is therefore regarded asunnecessary again to describe it in detail.

The various cooling arrangements andl oil pressure regulating and pressure surge cushioning devices disclosed in connection with any of the above embodiments may obviously be employed in connection with the others. Moreover any suitable fluid may be used as. hydraulic medium in substitution for the lubricating oil. And it is to be expressly understood that the invention is notlirnited to the specific embodiments shown or described, but may be used in various other ways,` and various modifications may be made to suit different requirements, and` that changes, substitutions, additions and omissions. may be made in the construction, arrangement and manner of adjustment and operation within the limits or scope of the invention as defined in the following claims.

Where the claims are directed to less than all the elements of the complete system disclosed, they are intended to cover possible uses of the recited elements in installations which lack the non-recited elements.

I claim:

l. For use with au internal combustion engine and the like having a valve mechanism including valve means actuated from the engine in timed relation therewith, the combination with valve timing control means, of engine speed responsive means, engine torque responsive means, and an operative connection for actuating said timing control means from said speed and torque responsive means.

2. In a valve mechanism for an internal combustion engine and the like having valve means actuated from the engine in timed relation therewith, the combination with valve timing control means, of a torque meter, and an operative connection for actuating said valve timing control means from said torque meter.

3. In a fuel feeding system for a combustion engine, the combination with valve means actuated by the engine in timed relation therewith for intermittently supplying fuel to the engine, of timing control means for said valve means, means responsive to changes of engine torque, and an operative connection between said last mentioned means and said timing control means for varying the timing of the valve means automatically.

4. For use with a combustion engine, a fuel feed system having reciprocable valve means for intermittently supplying full thereto, timing control means co-operating with said valve means for varying the timing of the intermittent fuel supply, means responsive to changes of engine torque, and an operative connection between said torque responsive means and said timing control means to actuate the latter automatically.

5. Engine valve mechanism having intake and exhaust valves and an actuating mechanism so arranged as to open the exhaust valve before the closing of the intake valve, first means for variably regulating the valve overlap; and means automatically actuated upon variations of engine torque for operating said first means.

6. Engine valve mechanism having admission and exhaust valves and an actuating mechanism for said valves so arranged as to obtain overlapping opening of said admission and exhaust valves, iirst means for variably regulating the valve overlap; and means automatically actuated upon changes of engine speed and torque for operating said first means.

7. Engine distribution system including valve means, a mechanism operating said valve means at every engine cycle, regulating means for varying the timing of said valve means, means responsive to changes of engine torque, and an operative connection between said torque responsive means and said timing reguiating means.

8. Engine distribution means including a valve actuated at every engine cycle, regulating means for varying the timing of said valve means, means responsive to changes of engine torque, means responsive to variations of engine speed, and an operative connection actuating said timing varying means from said torque and speed responsive means.

9. In an internal combustion engine, the combination in a fuel distribution system, of means cyclically actuated from the engine for intermittently admitting fuel thereto. timing regulating means for variably adjusting the timing of the intermittent admission of fuel with respect to the engine cycle, torque responsive means for sensing variations` inthe torque developed by the engine, and means for operatively connecting the torque responsive means tothe timing regulating means to vary the timing of the intermittent fuel admission as a preselected function of the engine torque.

l0. The combination defined in claim 9 further including engine speed responsive means and an operative connection between said speed and torque responsive means for varying the eiect of changes in engine torque at different engine speeds.

11. In a distribution system for an internal combustion engine, the combination with a rotary distribution member, of driving means for causing rotation of said member at a speed related to engine speed, regulating means for 14 varying the angular phase of said member with respect to the engine cycle to alter the distribution timing thereof, engine torque responsive means, and an operative connection for adjusting said regulating means from the torque 5 responsive means.

References Cited in the le of this patent UNITED STATES PATENTS Kittler Apr. 7, 1942

Images