US3739809A - Engine apparatus - Google Patents

Abstract

The fuel-per-stroke which fuel injection systems deliver to an engine is varied in accordance with engine speed to provide a desired fuel-air mixture ratio by means of a spring-loaded check valve responsive to momentary pressure impulses occurring in a pump chamber and operative to divert increasing fuel at increasing engine speeds. The check valve spring loading is varied simultaneously with the pump delivery setting in some embodiments, and valve operation is varied in accordance with engine speed and/or acceleration in some embodiments. The system is illustrated in connection with several reciprocating fuelinjection pumps and several rotary fuel-injection pumps.

Description

[ June 19, 1973 ENGINE APPARATUS Otmar M. Ulbing, R.D. No. l, Berkshire, NY.

[22] Filed: June 21, 1971 [21] Appl. No.: 155,114

Related US. Application Data [76] Inventor:

[52] US. Cl. 137/565.2, 417/255, 417/289,

[51 int. C1. F04b 49/00, F04b 39/00 [58] Field of Search 417/289, 303, 310,

[56] References Cited UNITED STATES PATENTS 2,117,512 5/1938 Scott 417/462 3,057,300 10/1962 Ulbing 137/565.l

3,120,814 2/1964 Mueller 417/310 Primary Examiner-William L. Freeh Attorney-Richard G. Stephens [5 7] ABSTRACT The fuel-per-stroke which fuel injection systems deliver to an engine is varied in accordance with engine speed to provide a desired fuel-air mixture ratio by means of a spring-loaded check valve responsive to momentary pressure impulses occurring in a pump chamber and operative to divert increasing fuel at increasing engine speeds. The check valve spring loading is varied simultaneously with the pump delivery setting in some embodiments, and valve operation is varied in accordance with engine speed and/or acceleration in some embodiments. The system is illustrated in connection with several reciprocating fuel-injection pumps and several rotary fuel-injection pumps.

28 Claims, 19 Drawing Figures Patented June 19, 1973 10 Sheets-Sheet 1 ni F INVENTOR.

I OTMARH-ULBING Way Patented June 19, 1973 10 Sheets-Sheet 2 Patented June 19, 1973 3,739,309

10 Sheets-Sheet 3 I MINIMUM DELIVERY III III MEDIUM MAXIMUM DELIVERY DELIVERY Patented June 19, 1973 10 Sheets-Sheet 4 NQK Patented June 19, 1973 3,739,809

10 Sheets-Sheet 5 Patented June 19, 1973 10 Sheets-Sheet 6 In 2 E F Patented June 19, 1973 3,739,809

10 Sheets-Sheet '7 a! c I d e f K r1 :1 FL FL FL 11 I I g h I I k J J TL 1 FL FL m I z 1 :12; b m c n: d I::! e :3 f I H I I 9 h J I: k

I o. b c d f G l l l I l l g h i k F -1 l 'lT [T m W l E 3 7 ha 2 a: D 'Z} q h E i C: j 2:: k 4:: ffl L1 I I J! L J I JTL -J W REV I PLATE 225:! 1:: I: L71: II': I: A 7- I *r- I W 9 0 IB O 2 /0 36 0 Patented June 19, 1973 10 Sheets-Sheet 8 ENGINE FIG. 30.

CO AP NOZZLE SUPPLY Patented June 19, 1973 3,739,809

10 Sheets-Sheet 9 Patented June 19, 1973 10 Sheets-Sheet 10 ENGINE APPARATUS This application is a division of my prior copending application Ser. No. 857,162 filed Sept. 11, 1969 now US. Pat. No. 3,614,944, and a continuation-in-part of my copending application Ser. No. 786,233 filed Dec. 23, 1968. This application relates to improvements to the apparatus shown in prior application Ser. No. 786,233. Certain principles and aspects of the present invention are readily applicabble, moreover, to internal combustion engine fuel systems other than those shown in my prior application.

It is desirable, after starting, and during normal warmed up operation, that the fuel-to-air ratio of the mixture burned by an engine either remain relatively constant or vary slightly in a desired sense over a wide range of speed and load conditions. It is generally preferred that the ratio either remain constant or vary slightly inversely with engine speed. Most internal combustion engines are controlled by a primary control which operates basically to control engine torque. In most carburetor-equipped engines the engine primary control controls torque by adjusting a throttle plate to adjust the amount of fuel-air mixture aspirated into an engine cylinder during a stroke. The amount of mixtureaspirated per stroke determines the amount of air drawn into the carburetor air intake per stroke, and because of the venturi principle upon whichthe carburetor operates, the amount of fuel mixed with air varies rather proportionally. Because carburetor air flow inherently controls carburetor fuel flow, the fuel-to-air ratio of the mixture inherently tends to remain somewhat constant over a wide range of engine speeds. Any unwanted variation can be corrected easily in a number of ways, such as by use of a metering needle valve moved in response to engine vacuum.

While fuel injection engine systems have a number of advantages over carburetor-equipped engine systems for a number of applications, fuel injection systems do not have the above-described automatic mixturemaintaining tendency of carburetor systems. In fuel injection systems engine torque is usually controlled by simultaneous variation of the amount of fuel pumped per stroke and the amount of air aspirated per stroke, with the engine torque control linked to both vary the amount of fuel pumped per stroke and to vary a throttle plate in the air intake duct. When a decrease in engine load causes an increase in engine speed at a given torque control setting, the increased air flow causes an increased pressure drop across the air intake structure, decreasing the amount of air aspirated per stroke, With the amount of air aspirated per stroke decreased and the amount of fuel pumped per stroke remaining the same, it will be seen that the fuel-to-air ratio disadvantageously tends to increase as enginespeed increases due to decreased load at a given torque controlsetting. While such a variation in fuel-air ratio is not regarded as a serious limitation in many applications, in certain other applications, and particularly in those where the engine frequently operates under widely-varying load conditions, it is desirable that fuel delivery vary inversely with engine speed, or directly with engine load, as well as with adjustment of the engine torque control, or primary control. The variation of fuel delivery so as to maintain a desired fuel-to-air relationship may be termed secondary control. it is a principal object of the invention to provide improved fuel injection systems having effective and reliable secondary control.

Secondary control, or automatic variation of fuel delivery with engine speed in order to keep the fuel-air mixture relatively constant, is known'in the prior art in connection with diesel engines and certain gasoline engines which use fuel injection. One prior art system is shown in US. Pat. No. 3,443,554, for example. However, those prior art systems of which I am aware require a centrifugal governor and/or other very complex and expensive mechanisms in order to provide secondary control. A very important object of the invention is to provide secondary control using much simpler and more economical apparatus which is easily constructed and highly reliable. Use of the present invention allows one to replace extremely complex and expensive mechanisms with merely a spring-loaded check valve.

In some prior art fuel injection systems themaximum pressures developed in the injection pump during an engine cycle vary appreciably with both engine speed and the fuel delivery setting of the basic torque control, while the maximum pressure developed in various other systems varies substantially solely with engine speed and is substantially independent of the fuel delivery setting of the basic torque control. As will be seen below, secondary control of some systems may be obtained solesly by use of a spring-loaded check valve having a constant spring load, and secondary control of the other prior systems may be obtained very simply by use of a spring-loaded check valve together with means for varying the spring load as a function of the fuel delivery setting of the basic torque control. Thus further objects of the invention are to provide effective secondary control for fuel injection systems of both types.

The basic principles of the present invention are readily applicalbe both to fuel injection systems used with four-cycle engines and those used with two-cycle engines, and readily applicable to both fuel injection systems used with a single cylinder and those used with multi-cylinder engines, including those having distributor arrangements to successively feed plural cylinders in a sequence, and further objects of the invention are to provide apparatus having such versatility.

Several basically-different types of variable-delivery reciprocating pumps are used in fuel injection systems, and another object of the invention is to provide secondary control arrangements which are useful with each of the different types of pumps. Several different types of variable-delivery rotary pumps are also used in fuel injection systems, and a further object of the invention is to provide a secondary control arrangement which may be used with one or more such pumps. The invention is applicable to any fuel-injection system in which a pulsating pump is used and/or in which a distributor is used, in such a manner that a peak pressure impulse varying in accordance with engine and pump speed occurs during a pumping cycle.

A further object of the invention is to provide a simple secondary control device which may also be arranged to act as an engine speed governor or limiter.

Yet another object of the invention is to provide a single secondary control device which also may be arranged to be responsive to engine acceleration and deceleration to vary the fuel-air ratio in a desired sense when the engine accelerates or decelerates. A further object of the present invention is to provide an improved form of a special type of rotary pump in which control of delivery is effected by limited angular adjustment of a substantially stationary member rather than by controlling the phase relationship of two rapidlyrotating components.

Other objects of the invention will in part be obvious and will, in part, appear hereinafter.

The invention accordingly comprises thefeatures of construction, combination of elements, and arrangement of parts, which will be exemplified in the constructions hereinafter set forth, and the scope of the invention will be indicated in the claims.

For a fuller understanding of the nature and objects of the invention reference should be had to the following detailed description taken in connection with the accompanying drawings, in which:

FIG. 1 is a cross-section view ofa reciprocating variable-delivery pump of the type shown in my copending application modified to incorporate secondary control in accordance with thepresent invention.

FIG. 1a shows a modification to the pump of FIG. 1 to provide a different type of oil pumping.

FIG. 1b is a porting diagram useful in understanding theoperation of the pump of FIG. 1.

FIGS. 2a and 2b are cross-section views of a rotary variable-delivery distributing pump of a basic type shown in my prior US. Pat. No. 3,057,300 modified to incorporate secondary control in accordance with the present invention. FIG. 2a is a view taken at lines 2a2a in FIG. 2b and FIG. 2b is a view taken at lines 2b2b in FIG. 2a.

FIG. 20 is a view of one plate part of the apparatus of FIGS. 2a and 2b.

FIG. 2d is a view of a portion of FIG. 2a.

FIG. 2e is a view of an adjustable metering plate part of the apparatus of FIGS. 2a and 2b.

FIG. 2f is a timing diagram useful in understanding the operation of the apparatus of FIGS. 2a and 2b.

FIG. 2g is a cross-section view taken at lines 2g2g in FIG. 2d.

FIG. 2h illustrates certain modifications which may be made to the device of FIGS. 2a and 2b.

FIGS. 3a, 3b and 3c are diagrammatic views useful in explaining the operation of the invention with each of three different types of reciprocating, variabledelivery, constant stroke-length fuel injection pumps.

FIG. 4 is a cross-section view of a modified form of rotary variable-delivery fuel injection pump and distributor.

FIG. 4a is a view of a rotatable distributor plate of the device of FIG. 4.

FIGS. 5a to Sc illustrate an alternative embodiment of the invention. FIG. 5a is a cross-section view through the pump, FIG. 5b is a view taken at lines Sb-Sb in FIG. 5a, and FIG. 50 is an exploded view in which the parts are shown isometrically FIG. 3a diagrammatically illustrates the application of the invention to a well-known form of fuel injection system using a variable-delivery reciprocating pump in which delivery is varied by varying the time or point during the stroke at which an inlet port is closed off. The engine crankshaft is mechanically connected to reciprocate piston P a fixed stroke distance s within'cylinder CY. Piston P is provided with an axial length which varies angularly around the piston, so that rotating piston P to various angular positions will vary the time or position during each stroke at which the piston covers inlet port IP. If port IP is closed off early during the stroke, greater delivery will result. During a rightward pumping stroke, fuel will be expelled through inlet port IP back to the supply tank until the piston blocks the port, and forward pumping will occur through delivery check valve DV to nozzle NO after the piston blocks the port until the end of the stroke. The angular position of piston P is controlled by the engine primary control, which is shown as comprising an accelerator pedal A. The primary control also varies air intake throttle plate TP via a cam or suitable linkage I to generally increase air flow as fuel quantity is increased. At a given setting of the primary control, an increase in engine speed due to a decrease in load will decrease the amount of air aspirated per stroke due to the restriction of the air intake structure S, thereby undesirably increasing the fuel-air mixture ratio.

As piston P of the injection pump travels rightwardly on a pumping stroke from the leftward limit position shown in FIG. 3a, fluid will initially be expelled out inlet port IP. If piston velocity were uniform, the pressure within the cylinder during the initial travel would tend to be a constant value dependent upon piston speed and the restriction of unblocked port IP. If piston P is instead reeiprocated with simple harmonic motion, or an approximation or modification thereof, as is usually the case in practical systems using cranks or eccentrics or the like, the linear veloctiy of the piston will instead vary approximately cosinusoidally, for example, from zero velocity at the leftward position shown to a maximum speed at mid-stroke, down to zero velocity at the end of the rightward stroke. The linear velocity, assuming simple harmonic motion, may be written as (taking mid-stroke as the origin):

v=s/2 w COS wt With piston velocity increasing during the initial portion of the stroke, the flow Q through port IP will similarly increase in direct proportion. FRom Torricellis theorem Q /A it is evident that the pressure drop across the port, and hence th pressure within the cylinder, will vary as the square of the flow Q, and vary inversely with the square of the area of the port, and hence the pressure will increase approximately in accordance with an w (cos cut) characteristic, where w is engine speed in radians per second. Thus the pressure existing in the cylinder when the piston edge reaches the inlet port will be higher at higher engine speeds, varying approximately with the square of engine speed. If the area A of port IP is large, the pressure built up in the cylinder prior to closure of the inlet port will still be very modest, however, and in many systems is small enough to be neglected.

If at a given engine speed the piston is rotated to close off port I? later during the stroke, so as to provide lesser delivery per stroke, but still prior to midstroke,

the piston will be seen to reach a greater linear velocity by the time it begins to close off port IP, and hence a greater pressure will exist when closure begins than at greater delivery settings. At any given engine speed, the maximum pressure would be developed if port IP is closed off approximately at midstroke, when the piston is at its maximum linear velocity. If piston P is rotated to provide port closure very late during the stroke, a lesser peak pressure will be developed due to the lesser piston velocity at the time of closure. Prior art systems of the type shown in FIG. 3a did not include the further check valve SV, and its presence should be ignored for the moment. i

As the piston begins to close off port 1?, th pressure further increases,'notonly due to the increase in piston lineanvelocity and increase in flow, (assuming port closure prior to midstroke) but also due to the decrease in unblocked port area irrespective of whether port closure occurs before or after midstroke. The pressure p rises in the cylinder as the port is closed off roughly in accordance with the following characteristic:

where A, is the initial open area of the port, d is' the axial width of the port, and k and k are constants. As the port is gradually closed off it becomes an infinite restriction, and hence as theport is closed off the pressure rises exponentially, theoretically toward an infinite value. However, when the pressure reaches a value determined by the spring loading of delivery check valve DV, the valve DV opens and the increase in pressure is thereafter limited/Thefrate, at which the pressure rises as the port is closed off depends not only upon the geometry of the port,-.but also upon the piston speed; The increase in piston speed not, only increases. the pressure dueto provision of increased flow, but al'soincreases the rate of pressure increase bymore quickly closing off the port, so that the rate of pressure-increase varies as a fairly high power of engine speed. The precise slope of the pressure characteristic will also depend, of' course, upon the shape of the inlet port as well as its general width, andthe shape of the piston edge.

which passes over the port to block the port-As the piston is rotated to decrease the delivery per stroke at a given engine speed, sothat the pistonhas a. greater velocity when it. c losesoff the port, the rate of pressure increase willbe seen to. increase. Since maximumlinear piston ve-loctiy, occurs at midstroke, the maximum peak pressur'e forlany given engine speed will occur 'when the portis closed off approximately at midstroke, when the pump is: adjusted to'pump approximately one-half its maximum. delivery perstroke.

The rapid increase in pressure as the portfisiclose'd off applies a sudden force to the body of valve' DV, accelerating it rightwardly against the force of the valve spring and providing a damped oscillation of thevalve body. The mass of the valve body, the valve spring, and the viscous resistance of the fuel to motion of the valve bodyafter the valve is opened will be seen to provide a mass-spring-damper second order system. The valve body, eventually returns to a steady-state position such that cylinder pressure balances the valve' spring: loading, and cylinder pressure remains substantially at the value determined by the delivery valve spring loading for the remainder of the rightward pumping stroke; The

motion of the check valve body required to allow maximum flow through the check valve is assumed to be small compared to the length of the check valve spring,

and hence the spring may-be assumed to apply a substantially constant force to the check valve body. Because the pressure drops quickly when the delivery check valve opens, the force applied to the valve body has the nature of a brief impulse, the amplitude of which varies as an exponential function of engine speed. Thus increased engine speed increases the amount which the valve body overshoots. After the overshoot, the check valve maintains cylinder pressure substantially at a value determined by the check valve spring loading. The pressure in the cylinder may increase slightly up to midstroke a's piston velocity increases and thereafter decrease somewhat as piston.- velocity decreases during the latter half of the rightward pumping stroke but no further sudden increase in pressure will occur during the pumping stroke. As mentioned above, the foregoing description of operation assumes that check valve SV is not present.

In accordance with the present invention, fuel delivery per stroke may be descreased with increasing engine speed by provision of the further secondary control check valve SV, which is responsive to pump cylinder pressure and operative to spill back increasing amounts of fuel to the supply as engine speed increases. The ratio between the mass of the body of a check valve to the spring force of the spring of the check valve may be termed the check valve time constant. Secondary control check valve SV is provided with a smaller time constant than that of delivery check valve DV. As the closure of inlet port I? causes the rapid increase in cylinder pressure, the pressure is applied simultaneously to both the delivery check valve and the secondary control check valve. The pressure temporarily rises above the steady-state delivery valve pressure setting due to the greater inertia or longer time constant of the delivery check valve, which delays its opening. During that temporary high pressure condition the secondary control check. valve SV opens, due to its lesser mass, despite its greater spring loading, .and opening of valve SV spills back fuel to the supplytank and limits the pressure developed in the cylinder. The amount which valve SV opens will be seen also to depend upon the peak pressure developed in the cylinder, and hence upon engine speed. As well as improving mixture ratio by spilling back some fuel, the quick The variation in volumetric efficiency, or air aspirated per stroke, with speed is ordinarily non-linear for most engines. Also, the variation of peak pump cylinder pressure with pump speed is non-linear, and the variation in the amount of fuel which a typical springloaded check valve will pass with a given pressure impulse applied to it is also non-linear. Furthermore, the

peak pump cylinder pressure occuring at a given engine speed varies in accordance with fuel delivery setting, as described above. Because of these varying relationships, it is sometimes difficult to provide a desired fuelair ratio characteristic over widely-varying load conditions if a fixed spring loading is used on the secondary control check valve. In accordance with a further feature of the invention, the loading on the secondary control check valve may be varied as a function of the primary control delivery setting, and in FIG. 3a cam C rotated by the primary control A is effective to vary the spring load on check valve SV.

If the piston in FIG. 3a is rotated to decrease fuel delivery by closing off the inlet port later during the first half of the stroke, the piston will have a greater velocity as it closes off the port, thereby increasing the slope of the pressure characteristic, aswill be apparent from expression (2), and thereby providing greater impulses to open the secondary check valve SV. Piston velocity decreases during the latter half of the pumping stroke. Thus maximum peak pressure for a given engine speed is developed if the inlet port is closed approximately at midstroke, which results when the pump is operating at roughly one-half of its capacity. Most engine systems require fuel delivery which varies from none or some small amount up to a maximum required for normal running although even greater delivery may be required for starting. Since minimum delivery requires inlet port closure very late in the stroke, a pump of the type shown in FIG. 3a ordinarily will operate over a range which varies from a minimum delivery condition involving port closure very near the end of the rightward stroke when piston velocity is low, up to maximum delivery condition involving port closure much earlier during the stroke when piston velocity is greater. If the maximum fuel required by the engine during running conditions is no more than half the maximum pump capacity, it will be seen that the peak pump cylinder pressure developed at a given engine speed will vary directly, though not linearly, with the fuel delivery setting over the entire running range of the engine.v Under such conditions, cam C will ordinarily provide a spring-loading to valve SV which generally increases as the delivery is increased. If the engine requires more fuel delivery than half the pump capacity so that inlet port closure must occur prior to midstroke, a plot of the peak pump pressure developed at a given engine speed versus pump delivery will be seen to slop downwardly at the highest delivery values. Under such an arrangement, cam C will ordinarily provide spring-loading which increases as .delivery is increased upto a given delivery value, after which cam C will provide decreasing spring-loading as delivery is further increased.

FIG. 3b diagrammatically illustrates a different for of reciprocating variable-delivery constant strokelength injection pump in which delivery is varied by varying the time during the stroke at which forward pumping is terminated, rather than varying the time at which it begins. The pistons P and AP are reciprocated by the engine with some approximation of simple harmonic motion. As piston P travels rightwardly on a pumping stroke, delivery commences substantially immediately through delivery check valve DV, and continues throughout the rightward pumping stroke until port TP of auxiliary piston AP registers with port SP of collar CO, at which time fuel is spilled back through hose H to the supply tank. The delivery check valve feeds a nozzle extending into the engine air intake structure in the same manner as in FIG. 3a. The piston P contains a bore and a conduit communicating with port T? of auxiliary piston AP. Collar CO is arranged to be axially adjustable relativeto auxiliary piston AP by means of the engine primary control, so that the time or position during the stroke at which forward pumping ceases may be varied to vary the quantity of fuel delivered. Inlet check valve IV admits fuel to the pump cylinder during the leftward return or suction stroke.

First consider the operation without the use of secondary control check valve SV. At the beginning of a rightward pumping stroke, piston speed begins at zero and increases cosinusoidally. Pressure builds up in the pump cylinder substantially immediately to a value greater than the steady-state spring loading of the delivery check valve, and then decreases to a value commensurate with the delivery check valve loading, as the delivery check valve DV opens. While the velocity of the piston P is minimum (zero) at the beginning of the stroke, the acceleration of the piston is then at its maximum value, and assuming simple harmonic motion2a=sl2 (1) sin cut. The maximum acceleration of piston P applies a maximum impulse to delivery check valve DV, and the magnitude of the impulse will be seen to vary as the square of engine speed. As the delivery check valve opens, the pressure in the cylinder drops markedly. The pressure then increases somewhat until midstroke (assuming collar CO is adjusted to provide delivery past midstroke) due to the increasing velocity of the piston and increased flow through valve DV, but the pressure does not ordinarily approach the initial peak pressure. When port TP reaches port SP the pressure drops suddenly and delivery valve DV closes. Inasmuch as the peak pressure occurs at the beginning of the stroke, irrespective of the adjustment of collar CO, it will be seen that variation of the delivery setting of collar CO by the engine primary control has no effect on the peak pressure developed within the cylinder.

In accordance with the invention, secondary control check valve SV is provided in FIG. 3b, again with'a smaller time constant than delivery valve DV, so that valve SV opens briefly during the pressure peak to spill back fuel to the supply, and it will be apparent that increasing engine speed causes greater impulses to valve SV, thereby spilling back more fuel. Because the magnitude of the pressure peaks does not tend to vary with the delivery setting, it is in general less necessary to use a cam to vary the spring loading on the valve SV in FIG. 3b. However, the use of such a cam, in the same manner as in FIG. 3a, allows one to more easily tailor the secondary control valve spill-back amount to a given volumetric efficiency versus speed characteristic, and

the use of such a cam with the pump of FIG. 3b is within the scope of the invention.

While FIG. 3b illustrates a reciprocating variabledelivery pump using a constant stroke length, its peak pressure characteristic is essentially the same as that of a number of reciprocating variable-delivery pumps in which the amount of fuel pumped per stroke is varied by varying the pump stroke length. In such pumps, the peak pressure ordinarily occurs at or near the beginning of the stroke, and the magnitude of the peak pressure does not vary appreciably with the fuel delivery or stroke length adjustment. It will be apparent that a secondary control check valve may be connected to the chamberof such a pump in thesame manner as with the pump of FIG. 3b, with the check valve spring loading being either varied or notvaried as a function of the primary control or stroke length setting.

FIG. 30 diagrammatically illustrates a third form of reciprocating variable delivery, constant stroke-length injection pump of a type shown in greater detail in FIG. 1 and also described in detail in my copending application Ser. No. 786,233. Piston P is reciprocated by the engine with some approximation of simple harmonic motion. A passageway within piston P communicates with the pump chamber and selectively communicates with inlet, port IP and outlet port OP. The passageway edge positions vary angularly about the piston so that rotation of the piston varies the time during a given stroke at which inlet port I? is closed 'off and the time at which outlet port OP is opened, thereby varying the amount of fluid pumped during a rightward pumping stroke. The engine primary control rotatably adjusts piston P to vary pump delivery rate. At the leftward position of the piston inlet port I? is fully opened, and at the rightward end of the pumping stroke outlet port OP is fully opened. The passageway geometry is arranged relative to the two ports so that outlet port OP always opens slightly before inlet port IP is completely closed off at any angular position of the piston. With inlet port IP closing as outlet port OP is opening, the maximum restriction to flow from the pump cylinder occurs during the overlap condition when both ports are slightly open.

Consider initially the operation of the pump of FIG. 3c without secondary control check valve SV. As the piston begins a rightward pumping stroke, fluid is expelled through fully open inlet port 1?, and the pressure within the pump cylinder remains low. As the inlet port begins to close off and the maximum flow restriction condition is approached, the pressure in the pump cylinder increases very rapidly, and then as the maximum restriction overlap condition is passed and outlet port is opened wider, the pressure decreases. The magnitude of the peak pressure developed in the cylinder will be seen to depend upon both engine speed, which determines the flow rate out of the pump cylinder, and upon the minimum total open area of the two ports when both are slightly open. As the piston is rotated to vary the delivery, it will be seen that the time during the stroke at which the maximum restriction condition occurs will vary, and if the same maximum restriction condition occurs at different piston velocities, which provide different flow rates from the cylinder, it may be seen that the peak pressure obtained will also vary with the engine primary control setting. If the same maximum restriction condition, i.e., same minimum open area during overlap, is made to occur for all delivery settings, the peak pressure at a given engine speed will be seen to be obtained if the maximum restriction condition occurs substantially at midstroke, when piston linear velocity is greatest, so that the peak pressure for a given engine speed will occur when the pump is adjusted to pump approximately one-half of its maximum delivery per stroke. I

In accordance with the invention, secondary control check valve SV is connected from the pump cylinder to spill fuel during the occurrence of the pressure peaks. One advantage of the pump of FIG. 30 over those of FIGS. 30! and 3b is that the delivery check valve DV maybe very lightly loaded, since delivery cannot begin until output port OP is opened, irrespective of pump speed and delivery setting. Furthermore, while the peak pressure impulse developed in the pumps of FIGS. 3a and 3b and applied to their secondary control check valves is limited by the opening of their delivery check valves, the peak pressure developed in the pump of FIG. 30 is substantially independent of the delivery check valve loading, and thus the secondary control check valve used in the arrangement of FIG. 3c need not have a shorter time constant than that of the delivery check valve or otherwise be adjusted relative to any other check valve.

The above description of FIG. 30 assumes that the same maximum restriction condition occurs during the overlap condition at all angular adjustments of the piston. By suitably shaping and/or slanting the edges of the parts relative to the piston passageway edges one can cause the area of the maximum restriction to vary as the piston is rotated to provide different delivery rates, and hence one can make the amplitude of pressure peaks occurring in the pump of FIG. 30 either more a function of, or less a function of, the delivery setting in whatever manner one chooses. If the minimum port area occurring during overlap is caused to increase somewhat with delivery setting up to one-half of pump capacity, thereby decreasing the maximum restriction with an increased delivery setting up to one-half pump capacity, (and thereafter to decrease with increased delivery if more than one-half pump capacity is used) the magnitude of the pressure impulses will tend to vary less with delivery setting. If the minimum port area during overlap is caused to vary roughly as a sine-squared function with the delivery setting, it will be seen that the magnitude of the pressure peaks occurring at a given engine speed can be made theoretically independent of the delivery setting, so that no variation in check valve spring loading with delivery setting is necessary. Because the pump of FIG. 3a requires that the inlet port be fully closed, providing an infinite restriction at all delivery settings, the magnitude of the pressure impulses occurring in such a pump varies markedly in accordance with the delivery setting, since the delivery setting determines the time during the stroke, and hence the piston velocity at the time the restriction is imposed. The pump of FIG. 30 (and FIG. 1), by not providing an infinite restriction, but inStead a controllable partial restriction, the minimum area of which can be made to vary with delivery setting, therefore has the marked advantage that the magnitude of the pressure impulses occurring at a given engine speed may be arranged to vary with delivery setting in accordance with any desired function, or if desired, arranged not to vary appreciably at all. i

It has been shown that while the peak pressure developed during a pumping stroke at a given engine speed varies with primary control setting with the pump of FIG. 3a, with maximum pressure being developed when this pump is pumping at roughly one-half its maximum capacity, that the peak pressure developed in the pump of FIG. 3b tends to be largely independent of the primary control setting, and that the peak pressure developed at a given engine speed in the pump of FIG. 3c may or may not vary appreciably with delivery setting, depending upon whether its port geometry is arranged to provide a minimum restriction area which varies with delivery setting. The amount of fuelspilled back by the secondary control check valve of any of the three systems of FIGS. 3a, 3b and 3c varies with the peak pressure impulse applied to the check valve in a manner dependent upon the check valve passage geometry, as well as upon its inertia and spring loading.

FIG. 1 illustrates in a cross-section view a form of injector pump disclosed and described in detail in my copending application Ser. No. 786,233, with certain modifications made thereto in accordance with the present invention. The pump is of the basic type described above in connection with FIG. 3c, .but shown adapted for two-cycle engine use to pump oil as well as fuel. The pump comprises a generally-cylindrical central casting 120 having a rear head 121 and a front head 122 bolted thereto by means of bolts (not shown), with a suitable gasket (not shown) preferably provided between each head and the central casting. Shaft 123 rotated by the engine crankshaft carries eccentric cam 127. Rotation of cam 127 reciprocates tappet 81, which is carried in bushing 82 with an O-ring seal 83a. The right end of tappet 81 bears against the left end of piston 130, which reciprocates within sleeve 129a. A spring 133, only a portion of which is shown, is inserted between head 122 and a right-end face of piston 130 and operates to return piston 130. A lower gear sector 83 pinned to piston 130 is engaged by upper gear sector 84 pinned to control shaft 131, so that rotation of shaft 131 angularly positions piston 130. Shaft 131 is angularly positioned by accelerator pedal or throttle control 103 via arm 104 and a suitable mechanical linkage shown merely as a dashed line. Upper gear sector 84 is axially wider than lower gear sector 83 so that the gear sectors remain enmeshed as sector 83 reciprocates with piston 130.

Oil is supplied from an oil supply-tank (not shown) to chamber 128 via a check valve (not shown) and a pipe connection made at 128a on the side of central casting 120. Fuel is supplied to chamber 164 from fuel tank 146 via conduit 1450. Oil and fuel inlet ports are provided in sleeve 129a at 134 and 136, respectively, and oil and fuel outlet ports are provided at 135 and 137. Oil piston 161 is urged rightwardly against front head 122 by inner coil spring 162. Holes drilled in main casting 120 at 142a and 147a connect the outlet ports with longitudinally-extending passages in which check valves 150 and 151 are located, and plugs 1420, 147c close the ends of passages 142 and 147a. Check valves 150 and 151 at the outlet side of the injector pump each communicate with mixing chamber 143 provided in front head 122. Two V-shaped grooves are milled across the outer periphery of piston 30 as shown by dashed lines at 138 and 139. The bottom of V-groove 138 communicates with oil chamber 155 inside piston 130, and the bottom of V-groove 139 communicates with fuel chamber 160 situated to the right of piston 130 and partially within piston 130. At various axial positions of piston 130 V-groove 138 connects chamber 155 to only chamber 128 via oil inlet port 134, or to both inlet chamber 128 via inlet port 134 and to mixing chamber 143 via outlet port 135 and check valve 150, or to only mixing chamber 143 via outlet port 135 and check valve 150. At corresponding axial positions of piston 130, V-groove 139 connects fuel chamber 160 to only chamber 164 via inlet port 136, or to both chamber 164 via inlet port 136 and mixing chamber 143 via outlet port 137 and check valve 151, or to only mixing chamber 143 via outlet port 137 and check valve 151. Inlet ports 134 and 136 and outlet ports 135 and 137 each comprise an opening which extends partially around sleeve 129a, with each such slot having a uniform dimension measured in the axial direction of sleeve 1290.

The cutting of V-shaped grooves on the periphery of cylindrical piston 130 gives the grooves a width which varies with the angular position of the groove around the piston. As is described in greater detail in my copending application Ser. No. 786,233, varying the angular position of the piston within sleeve 129a by means of control shaft 131 varies the time during a given piston stroke at which the V-grooves will communicate with the outlet ports and the time at which the V-grooves will be cut off from the inlet ports, and hence determines the amount of fuel and oil which the pump will pass to the mixing chamber during the piston stroke.

Piston 130 is shown at its leftmost position in FIG. 1. As piston 130 is urged rightwardly on a pumping stroke, at the beginning of the stroke V-groove 138 connects oil piston chamber 155 via inlet port 134 to chamber 128 so that oil within chamber 155 is expelled from chamber 155 back into chamber 128, and V groove 139 connects fuel chamber 160 via inlet port 136 to fuel chamber 164, so that fuel is expelled from chamber 160 back into chamber 164. At an intermediate time during the stroke determined by the angular position of piston 130, the V-grooves first reach and unblock outlet ports and 137 and then move out of communication with and block inlet ports 134 and 136. Provision of such an overlap condition with the outlet ports always slightly opening before the inlet ports are fully closed prevents damage due to fluid blockage. Thereafter during the rightward pumping stroke, as the inlet ports fully close and the outlet ports increasingly open, oil is expelled from chamber via outlet port 135, and fuel is expelled from fuel chamber via outlet port 137, and the fuel and oil mix in mixing chamber 143. The mixing chamber connects to a nozzle (not shown) which injects the fuel-oil mixture into the engine air intake duct. As mentioned in my prior application, the fuel and oil are not mixed in a mixing chamber in some applications, and instead, only the fuel is piped to the nozzle and the oil is pumped to various oil holes at desired lubrication points within the engine.

The basic pump of FIG. 1 and FIG. 30 differs markedly from many somewhat similar prior art fuel metering pumps in that an inlet port is closed and a separate outlet port is opened during a pumping stroke, while the prior art generally (e.g. FIGS. 3a and 3b) has left each pump chamber in constant communication with an outlet check valve during the entire pumping stroke, so that forward pumpingpast a prior art check valve occurs either immediately (FIG. 3b) or as soon as the inlet port is closed off to prevent return pumping (FIG. 3a). If the fluid supply has positive pressure, the check valve in such prior systems must be loaded to at least the same pressure in order to prevent forward pumping prior to complete closure of the inlet port. And even if the fluid supply is not pressurized, the pressure in the prior artpump chambers necessarily builds up prior to complete closure of their inlet ports, in amounts dependent upon pump speed and dependent upon the amount of restriction to return flow between the pump chamber and the fluid supply, with the amountof said restriction increasing from a basic amount to complete blockage as the inlet port is gradually closed off. If for ward pumping is not to occur prior to complete closure of the inlet port, the check valve in the prior systems must be loaded to the highest such pressure which may occur prior to inlet port closure. The heavier check valve loading necessarily results in higher pressures in the pump chamber, thereby requiring a more precise piston-cylinder fit. In the pump of FIGS. 1 and 3a, forward pumping cannot occur prior to opening of an outlet port, irrespective of whether the supply is pressurized, and hence the instant at which forward pumping begins during a pumping stroke remains substantially independent of pump speed and outlet check valve loading, making the quantity of fluid delivered per stroke similarly independent of pump speed and check valve setting.

In accordance with the embodiment of the present invention illustrated in FIG. 1, fuel chamber 160 is connected to fuel inlet chamber 164 via a spring-loaded check valve 163, the spring loading of which is shown made variable as a function of control shaft 131 position, by means of cam 131a carried on control shaft 131. Rotation of control shaft 131, as by means of accelerator pedal 103 and arm 104, so as to rotate piston 130 to increase oil and fuel flow rates causes cam 131a to vary the spring loading on check valve 163. The precise shape of cam 131a will depend upon the desired variation of fuel-air ratio, the variation in air flow with engine speed due to the engine air intake structure, the variation of pump cylinder peak pressure with engine speed, the variation of pump cylinder pressure with delivery setting, and the variation in the amount of fuel spilled back through check valve 163 with peak pressure, all of which determine the variation in the amount of fuel spilled back for a given engine speed with a given primary control delivery setting. In some embodiments of the invention, the spring loading of check valve 163 need not be varied as a function of throttle position. In those embodiments can 131a may be eliminated and check valve 163 held in position with a fixed spring loading by a plug in head 122. The passageway which includes a check valve 163 extends generally in a direction so as to intersect shaft 131 if cam 131a is used. If no cam is used it will be apparent that the passageway may extend out radially in another direction, such as perpendicularly to the plane of FIG. 1.

In the pump of FIG. 1 the inlet and outlet ports are spaced relative to their respective V-grooves so that maximum restriction to flow from each V-groove occurs during the intermediate or overlap interval when each V-groove slightly communicates with both its inlet port and its outlet port. Therefore, the maximum pressure which occurs in pump cylinder 160 during a pumping stroke occurs during that intermediate or overlap interval when both inlet port 136 and outlet vary with pump piston speed. The pressure in chamber 160 will be seen to drop from its maximum value as piston 130 thereafter continues to travel rightwardly and outlet port 137 increasingly unblocked.

As was explained above in connection with FIG. 3c, the maximum peak pressure developed in the pump cylinder for any given engine speed tends to occur if the maximum restriction condition when both inlet an outlet ports are slightly open occurs when the piston has maximum linear velocity. Maximum piston velocity usually occurs somewhere near midstroke if an approximation of simple harmonic motion is used to reciprocate the piston, and adjustment of the pump to cause the two ports to overlap around the midstroke causes the pump to operate at approximately one-half its maximum capacity. If one-half or less of the pump maximum capacityis sufficient to supply the maximum fuel requirements of the engine, the overlap will occur during the last half of the pumping stroke, and if the shape and spacing of ports 136 and 137 and V-groove 139 provide the same minimum area restriction as piston 130 is rotated to give different delivery rates, increasing the primary control setting to call for increased delivery will increase the peak pressure developed for a given engine speed and tend to increase the amount of fuel spilled back by the secondary control check valve 163, and in such an arrangement cam 163 may be shaped to provide an increase in check valve spring loading as the primary control setting is adjusted to provide greater fuel flow. If, on the other hand, the maximum fuel requirements of the engine require more than one-half pump capacity, so that forward pumping is sometimes required during the first half of the stroke, and the port and V-groove geometry again provides the same minimum area restriction at different angular positions of piston 130, the cam may be shaped to increase check valve spring loading until the primary control is adjusted to the midstroke overlap condition, and thereafter to decrease the spring loading as greater amounts of fuel are called for. However, if rotation of piston 130 is arranged to vary the minimum area of the maximum restriction which occurs during the overlap condition, the maximum pressure developed in cylinder 160 can be made to vary directly with delivery setting, or not to vary appreciable with delivery setting, or even to vary inversely with delivery setting, if desired. If the maximum pressure does not vary apprecialby with delivery setting, it will be apparent that variation of the spring loading on check valve 163 becomes unnecessary.

FIG. 1b contains three unrolled or developed views illustrating the geometry of V-groove 139 relative to ports 136 and 137. Angular adjustment of piston 130 to provide different delivery rates amounts to vertical displacement of the V-groove in FIG. lb relative to ports. On each pumping stroke V-groove 139 moves rightwardly relative to the ports from a beginning position in which the V-groove is centered on the inlet port port 137 are both only slightly open, so that maximum restriction to flow from chamber is provided, and the magnitude of the maximum pressure will be seento 136. V-groove 139 is shown at I in a minimum delivery position at the time during the overlap condition when p it least registers with inlet port 136 and outlet port 137, at II in a medium delivery position at the time during the overlap condition when it least registers with the ports, and at III in a maximum delivery position at the time when it least registers with the ports. It will be seen that the minimum overlap area varies from a small area in I, to a larger area in II, and then to a smaller area at III. Times t,, t and 2 indicate the times after the beginning of the pumping strokeat which the maximum restriction occurs under the three different delivery conditions. Thus it will be seen that the amount of maximum restriction varies from a minimum at low delivery rates up to a maximum at approximately one-half capacity, down to a minimum at maximum delivery. Since piston speed at the time of the overlap condition varies in approximately the same manner, it will be apparent that the variation in restriction may be used to offset the variation in piston speed at the time of overlap, so that the magnitude of the pressure impulses developed at a given engine speed tends to be largely independent of the pump delivery setting.

If cam 131a is eliminated and a constant spring load is used on check valve 163, and if the port geometry provides the same minimum restriction at different delivery settings, the amount of fuel spilled back through the check valve at a given engine speed will increase with the primary control delivery setting as the primarycontrol is varied from minimum flow to one-half pump capacity, thereby leaning out the fuel-air ratio, and as the primary control delivery setting is further advanced at the same engine speed to provide greater flow than one-half pump capacity, the amount of fuel spilled back through the check valve will decrease, thereby providing an increasingly-enriched mixture at increasing delivery settings.

In the arrangement shown in FIG. 1, wherein increasing fuel spill-back occurs at increasing engine speeds due to light load conditions but no comparable oil spillback occurs it will be seen that the amount of oil pumped per stroke remains substantially constant, thereby providing larger oil-to-fuel and oil-to-air ratios during higher speed-lighter load conditions. Such operation is wholly satisfactory for many two-cycle engine applications, and particularly in those two-cycle engine applications where the oil is not mixed with the fuel but instead pumped to various lubrication joints within the engine.

If desired, the oil-to-fuel and oil-to-air ratios may be tailored by providing a secondary control oil check valve in similar fashion to spill back oil in amounts varying with engine speed.

While FIG. 1 illustrates a system which dispenses metered amounts of oil as well as fuel, such as is used with two-cycle engine systems, it is important to recognize that the invention is in no way restricted to fuel injection systems which dispense two fluids, and is quite as applicable to four-cycle engine systems wherein oil is not injected into theengine.

While the mixing and metering pump of FIG. 1 uses separate oil and fuel pistons (161 and 130) to pump oil and fuel with a desired ratio, an alternative embodiment shown in FIG. 1a dispenses with the need for a separate oil piston, and the need for V-groove 138 on piston 130 and the need for oil inlet and outlet ports 134 and 135 in sleeve 129a. In FIG. 1a oil is supplied to oil chamber 128 via an inlet conduit 1280 which carries duckbill check valve 601. As cam 127 moves fuel piston 130 on a rightward fuel-pumping stroke, thereby increasing the volume of chamber 128, oil is drawn into chamber 128 through check valve 601. As spring 133 moves piston 130 leftwardly, thereby decreasing the volume of chamber 128, oil is expelled past oil outlet check valve 150 to mixing chamber 143. The amount of oil which is drawn into chamber 128 during a rightward stroke and dispensed to the mixing chamber on the piston return stroke depends upon the cross-sectional area of piston times the length of the pistonstroke, less the cross-sectional area of tappet 81 times the same stroke length, since tappet 81 increasingly enters chamber 128 from bushing 82 as piston 130 increasingly leaves chamber 128. If tappet 81 is very slightly less in diameter than piston 130, very little oil will be pumped comparedto the amount of fuel pumped. It will be seen that a constant amount of oil will be pumped per stroke, irrespective of the adjustment of control shaft 131. In a variety of systems, and in particular those which drive constant loads, it is considered unnecessary to maintain a constant fuel-oil mixture ratio. The pump in FIG. 1a is shown without the cam 131a and check valve 163 utilized in FIG. 1 to provide secondary control, and such a feature obviously can be added to FIG. la, if desired.

The pressure-responsive secondary control concept ofthe present invention is not limited to use with reciprocating fuel injection pumps, and is also applicable to rotary fuel injection pumps of the type which incorporate distributors to distribute fuel to different engine cylinders. FIGS. 2a and 2b illustrate an application of the invention to a rotary fuel injection pump and distributor device of a basic type shown in my prior U.S. Pat. No. 3,057,300.

. The pump includes a main casting 201 and a head 202 bolted to casting 201 by means of bolts 203,203. Main casting 201 includes a cylindrical bore along axis x-x having three different diameters indicated at 206a, 206b and 2060. Drive shaft 207 extends through the bore, being journalled in portion 206a of the bore by means of bearings 208a, 208b. Seal retainer washer 210 and seal 211 seal the outer end of shaft 207, and ring 212 carrying seal 213 and rubber O-ring 214 seal shaft 207 adjacent bearing 208a. Aligning pin 201a seats in bores in casting 201 and head 202, and passes through slots in ring 234 and plate 225, thereby angularly fixing these parts relative to each other. Ring 212 is stationary, and earn 227 is attached to shaft 207 and rotatable therewith. Plate 232 is capable of limited angular adjustment about axis xx by means of control rod 204, which is reciprocated by adjustments of the engine primary control (not shown). Spacer ring 246 having a slightly greater axial thickness than plate 232 surrounds plate 232 and is angularly held by pin 201a. Provision of spacer ring 246 transmits the force of disc spring 221 from ring 234 to head 202, so that plate 232 is not clamped tightly between head 202 and ring 234 and can be angularly adjusted easily.

A passageway 216a, 216b in casting 201 connects to the fuel supply (not shown), thereby admitting fuel to a ring-shaped chamber 217 formed by an annular groove around the external periphery of ring 212. A plurality of passages 218,218 extend inwardly and axially in ring 212 to permit fuel to flow from chamber 217 through holes in ring 220. A dome-shaped spring cap or disc spring 221 retaining O- rings 222 and 223 urges plate 220 and ring 212 rightwardly in portion 2061) of the cylindrical bore. Stationary plate 225, which is shown in detail in FIG. 20 is mounted against tha back of disc spring 221. Plate 225 is provided with an oversize central bore greater than the diameter of shaft 207, provided with slot 224 to accommodate aligning pin 201a, and provided with six holes 225a225fspaced in a circle 60 from each other. The

embodiment shown is designed for use with a sixcylinder engine, and other hole arrangements are provided in plate 225 for other types of engines. Four holes arranged at 90 from each other would be used with a four-cylinder engine, for example.

The oversize central bores in disc spring 221 and fixed plate 225 permit fuel flow from chamber 217 through the holes in ring 212 and plate 220 to a ringshaped chamber 226 formed by a cylindrical recess in rotatable cam 227, which is shown in greater detail in FIG. 2d. Cam 227 (FIG. 2d) includes two portions 227a, 227C of slightly different radius, with transition slopes 227b, 227d between the two portions. The two transistion slopes are located 180 around the cam from each other. The cam portion 227a of greater radius may be termed the cam lobe. As cam 227 rotates clockwise as viewed in FIG. 2d, slope 227b acts as the leading edge of lobe portion 227a, and slope 227d acts as the trailing edge of thelobe. Inlet passage 228 extends radially within cam 227 from chamber 226 and opens on the side of the cam at trailing edge 227d, and hence it will be seen that trailing edge 227d is in constant communication with the fuel supply. Passage 230 extends inwardly within cam 227 from leading edge 227b to where it intersects outlet passage 231, an elongated radially extending slot which also extends axially through the cam. With cam 227 mounted adjacent fixed plate 225, as shown in FIG. 2b, it will be seen that cam outlet slot 231 will successively register with individual ones of the six holes 2250-225 f (FIG. 2c) in plate 225, and will not register with any hole in plate 225 at an intermediate angular position between a pair of holes in plate 225.

As seen in FIGS. 2b and 2d, non-rotatable ring 234 surrounding cam 227 contains two circular holes 234a, 234b each opening into its circular central bore 234s, and two partially-circular crescent-shaped cam followers 235, 236 having the same axial length as ring 234 and cam 227 seat within holes 234a and 234b, respectively. Compression spring 237 carried in a bore in ring 234 urges follower 235 counterclockwise in recess 234a, and compression spring 238 similarly carried in ring 234 urges follower 236 counterclockwise in recess 234b, and hence edges 235a and 236a of the cam followers seat against the periphery of cam 227.

With cam 227 in the position shown in FIG. 2d, it will be seen that leading edge 227b and outlet slot 231 connect to a relatively large chamber bounded by edge 2350 of follower 235 and edge 236a of follower 236, and that trailing edge 227d and inlet passage 228 of the cam connect to a relatively small chamber bounded by edge 236a of follower 236 and edge 235a of follower 235. The chamber containing trailing edge 227d will be smaller than the chamber containing leading edge 227b because of the greater size of cam lobe portion 2270 as compared to cam recess portion 2270. As cam 227 rotates slightly more than 180 from the position shown, thereby moving the cam lobe 227a to decrease the size of the initially larger upper chamber, it will be seen that fuel will be expelled through passage 230 and slot 231, and as cam lobe 227. moves out'of th initially smaller lower chamber, it will be seen that fuel will be sucked into that chamber through inlet passage 228.

When cam 227 has rotated slightly more than 180 from the position shown in FIG. 2d, leading edge or rise 227b of the cam lobe will rotate follower 236 clockwise against the force of compression spring 238, and trailing edge 227d will release follower 235, allowing its compression spring 237, to rotate follower 235 counterclockwise. Such movement of the cam followers, so that their edges 235b, 236b now seat against the cam, will be seen to re-establish a relatively large chamber in front of leading edge 227b and a relatively small chamber behind trailing edge 227d, so that the next half-revolution of the cam again expels fuel out through outlet passages 230 and 231 and again draws in fuel through inlet passage 228. Thus as drive shaft 207 continuously rotates cam 227, fuel is constantly drawn into passage 228 and expelled through slot 231. The pressure of the fuel in outlet slot 231 will be seen to vary as a function of pump speed. Also, it will be seen that the flow through passage 231 will be substantially constant throughout a complete revolution of the pump cam, except for momentary decreases twice during each. revolution when cam followers 235, 236 are rotated. The cam followers are preferably located at an angular position around ring 234 so that the followers rotate at two times when cam slot 231 registers with a particular two of the six holes in plate 225, and then rotation of the cam followers will occur at a low differential pressure condition, when both cam inlet passage 228 and cam outlet passage 231 are connected to the supply.

As cam 227 rotates through a single revolution, slot 231 will be seen to pump fuel back to the supply through successive ones of the six holes in plate 225, and forward pumping out to an engine cylinder nozzle can occur only when slot 231 lies in between a pair of holes and does not register with any holes in plate 225. When cam slot 231 does not register with a hole in plate 225, fluid is pumped out of cam slot 231 leftwardly as viewed in FIG. 2b, and as fuel is pumped leftwardly out of slot 231, the proportion of the time it flows to an injection nozzle to the time it is returned to the supply, is governed by the adjustment of adjustable metering plate 232. Adustable metering plate 232, as best seen in FIG. 2e, contains a central bore 232m to slidingly carry plate 232 on drive shaft 207 and allow limited angular rotation of plate 232 about axis x-x. Plate 232 is shown in phantom in FIG. 2a in place relative to shaft 207. Control rod 204 passing through bore 241 in casting 201 and fitted with a seal 242 is provided with a yoke and engaging plate 232 by means of cotter pin 243. As control rod 204 is adjusted by means of the engine primary control, plate 232 is angularly positioned about axis x x relative to the holes in plate 225. Plate 232 is provided with six return pumping holes a through f arranged in a circle at one radius from axis x--x, and six forward pumping holes 3 through I arranged in another circle at a different radius, with the holes of the two circles angularly staggered, or out-ofphase' with each other as shown. Outlet slot 231 of cam 227 is provided with a length so that it may partially register with holes of bOth circles as the cam rotates. The position of outlet slot 231 relative to plate 232 at one angular'position of drive shaft 207 is shown in dashed lines at 231' in FIG. 2e. The width of cam slot 231 is chosen relative to the size and spacing of the holes in metering plate 232 so that slot 231 always registers at least slightly with either a return pumping hole of the inner circle or a forward pumping hole of the outer circle, and so that there is a slight overlap as the slot passes from a hole in one circle to a hole in the other circle. As cam 227 rotates the slot 231 will be

Claims (28)

1. A fluid pump, comprising, in combination: means defining a pump chamber having return and delivery ports; movable means within said chamber for pressurizing fluid within said chamber; means for supplying fluid to said chamber; means for cyclically opening and closing said ports in synchronism with the motion of said movable means to provide three distinct conditions including a first condition in which said chamber communicates solely with a first said port, a second condition in which said chamber communicates decreasingly with said first of said ports and increasingly with the other of said ports and a third condition in which said chamber communicates solely with said other of said ports, said ports being arranged so that the total area of said ports communicating with said chamber reaches a minimum during said second condition and the pressure in said chamber reaches a maximum during said second condition; check valve means connected to release fluid from said chamber to said means for supplying fluid when said pressure exceeds a predetermined vaLue; and an output conduit connecting said delivery port to a utilization device.

2. A pump according to claim 1 in which said means defining said pump chamber comprises a first plate with a recess having an open end and a second plate mounted adjacent said first plate, said second plate containing said return and delivery ports, and in which said means for cyclically opening and closing said ports comprises means for providing a relative rotation between said first and second plates so that said open end of said recess cyclically communicates with said return and delivery ports.

3. A pump according to claim 2 in which said check valve means comprises a spring-loaded check valve mounted in a passageway within said first plate, said passageway communicating with said recess and said check valve being adapted to move along a valve-operating axis, said valve-operating axis having a radial component of direction, whereby centrifugal force caused by rotation of said first plate varies the opening pressure of said check valve.

4. A pump according to claim 2 in which said check valve means comprises a spring-loaded check valve mounted in a passageway within said first plate, said passageway communicating with said recess and said check valve being adapted to move along a valve-operating axis, said valve-operating axis having a tangential component of direction, whereby angular acceleration of said first plate varies the opening pressure of said check valve.

5. A rotary pump comprising a first member defining a cylindrical bore; a rotor disc mounted to rotate within said cylindrical bore and having a periphery provided with two portions of different radius; a pair of follower means rotatably carried in said first member and spring-biased against said rotor disc to selectively divide the space of said bore surrounding said rotor into chambers which alternately increase and decrease in volume as said rotor disc is rotated, a first passage in said rotor disc connecting the chamber which is expanding to a fluid source, and a second passage in said rotor disc connecting the chamber which is contracting to outlet port means extending between opposite sides of said rotor disc, whereby rotation of said rotor disc creates fluid pressure in said outlet port means of said rotor disc; a first stationary plate engaging one side of said rotor disc, said plate having at least one opening connected to said fluid source, said outlet port means of said rotor disc being arranged to periodically communicate with said opening as said disc is rotated; a second plate engaging the other side of said rotor disc and mounted for limited angular adjustment, said second plate having at least one return opening communicating with said fluid source and at least one delivery opening communicating with a utilization device, said outlet port means being arranged to periodically communicate with said openings in said second plate as said disc is rotated, whereby pressure created in said outlet port means may provide fluid flow through one or the other of said openings of said second plate during time periods when said outlet port means does not communicate with said openings in said first stationary plate, and the fluid flow occurring during said time periods is selectively proportioned between said openings of said second plate by the angular adjustment of said second plate, thereby controlling the delivery of fluid to said utilization device.

6. A pump according to claim 5 wherein said openings in said plates are angularly spaced around the axis of rotation of said rotor disc so that said outlet port means successively communicates with said openings in the following sequence: first with said opening in said first plate, next with said delivery opening in said second plate and thirdly with said return opening in said second plate, whereby delivery of fluid to said utilization device begins at a predetermined angular position of said rotor disc and terminates at varying angular positions dependent upon adjusTment of said second plate.

7. A pump according to claim 5 wherein said openings in said plates are angularly spaced around the axis of rotation of said rotor disc so that said outlet port means successively communicates with said openings in the following sequence: first with said opening in said first plate, next with said return opening in said second plate and thirdly with said delivery opening in said second plate, whereby delivery of fluid to said utilization device begins at varying angular positions of said rotor disc dependent upon adjustment of said second plate and terminates at a predetermined angular position of said rotor disc.

8. A pump According to claim 5 wherein the sizes and spacings of said openings in said plates are arranged in relation to said outlet port means so that said outlet port means communicates with at least one of said opeings at all angular positions of said rotor disc.

9. Rotary distributor pump apparatus for supplying fluid successively and equally in controlled amounts to a plurality of output conduits, comprising, in combination: pump means having means defining a bore and a rotor adapted to rotate in said bore about an axis; a fluid source connected to said pump means, said rotor including an outlet port radially displaced from said axis to sweep in a circular path as said rotor is rotated, rotation of said rotor creating fluid pressure at said outlet port; stationary plate means having a plurality of openings spaced apart in a circular path to successively communicate with said outlet port as said rotor is rotated, each of said openings being connected to said fluid source; adjustable plate means capable of limited angular adjustment and having a plurality return openings spaced apart in a first circular arrangement to successively communicate with said outlet port as said rotor is rotated and a plurality of delivery openings spaced apart in a second circular arrangement to successively communicate with said outlet port as said rotor is rotated, each of said return openings being connected to said fluid source and each of said delivery openings being connected to a respective one of said plurality of output conduits.

10. Apparatus according to claim 9 in which said outlet port is arranged to communicate with at least one of said openings at all angular positions of said rotor.

11. Apparatus according to claim 9 in which said rotor is generally cylindrical and provided with two end faces, said outlet port extending between said two end faces, said stationary plate means being disposed against one of said end faces and said adjustable means being disposed against the other of said end faces.

12. Apparatus according to claim 9 having a passage in said rotor connecting said outlet port to said fluid source, said passage containing a spring-loaded check valve.

13. Apparatus according to claim 12 wherein the portion of said passage containing said check valve extends at least partially radially in said rotor, whereby centrifugal force affects the opening pressure of said check valve.

14. Apparatus according to claim 12 wherein the portion of said passage containing said check valve extends at least partially non-radially in said rotor, whereby angular acceleration of said rotor affects the opening pressure of said check valve.

15. Appparatus according to claim 9 wherein said means defining said bore includes a pair of passages connected between said bore and said fluid source, each of said passages containing a spring-loaded check valve.

16. Apparatus according to claim 15 having spring means acting between said check valves to apply equal spring loading to said check valves.

17. Apparatus according to claim 16 having cam means for varying said spring means to vary the spring loading applied to said check valves equally.

18. Rotary distributor pump apparatus for supplying fluid successively and equally in controlled amounts to a plurality of output conduits, comprising, in combination: pump means having means Defining a bore and a rotor adapted to rotate in said bore about an axis; a fluid source connected to said pump means, said rotor including an outlet port radially displaced from said axis to sweep in a circular path as said rotor is rotated, rotation of said rotor creating fluid pressure at said outlet port; stationary plate means having a plurality of return openings spaced apart in a circular path to successively communicate with said outlet port as said rotor is rotated, said stationary plate means also having a plurality of delivery openings spaced apart in a circular path to successively communicate with said outlet port as said rotor is rotated, said openings being phased so that said outlet port alternately communicates with return openings and delivery openings, said return openings being connected to said fluid source; a distributor plate arranged to rotate in fixed phase relationship with said rotor, said distributor plate having port means radially displaced from said axis to sweep in a circular path as said distributor plate is rotated, said port means of said distributor being phased relative to said outlet port of said rotor to communicate with individual ones of said delivery openings when said outlet port communicates with said ones of said delivery openings, respectively; and adjustable plate means capable of limited angular adjustment, said adjustable plate means having a plurality of return openings spaced apart in a first circular arrangement to successively communicate with said port means of said distributor plate as said distributor plate is rotated and a plurality of delivery openings spaced apart in a second circular arrangement to successively communicate with said port means of said distributor plate as said distributor plate is rotated, each of said return openings of said adjustable plate means being connected to said fluid source and each of said delivery openings of said adjustable plate means being connected to a respective one of said plurality of output conduits.

19. Apparatus according to claim 18 in which said rotor is generally cylindrical and provided with two end faces,said outlet port means extending between said two end faces, said stationary plate means comprising a first plate disposed against one of said end faces and containing said return openings of said stationary plate means and a second plate disposed against the other of said end faces and containing said delivery openings of said stationary plate means.

20. Apparatus according to claim 18 in which said outlet port of said rotor is arranged to communicate with at least one of said openings of said stationary plate means at all angular positions of said rotor.

21. Apparatus according to claim 18 having a passage in said rotor connecting said outlet port to said fluid source, said passage containing a spring-loaded check valve.

22. Apparatus according to claim 18 wherein said means defining said bore includes a pair of follower means spring-biased against said rotor to selectively divide the space of said bore surrounding said rotor into chambers which alternately increase and decrease in volume as said rotor is rotated, said bore being cylindrical and said rotor having a periphery provided with two portions of different radius, said portions of different radius being operative to control said follower means as said rotor is rotated.

23. Apparatus according to claim 18 wherein said bore is provided with two internal surface portions of different radius and said rotor carries a pair of follower means spring-biased to selectively divide the space of said bore surrounding said rotor into chambers which alternately increase and decrease in volume as said rotor is rotated, said internal surface portions of different radius being operative to control said follower means as said rotor is rotated.

24. Apparatus according to claim 18 Wherein said openings in said plate means are angularly spaced around said axis so as to provide fluid flow from said outlet port of said roTor in the following sequence: (1) through one of said return openings of said stationary plate means, (2) through one of said delivery openings of said stationary plate means, through said port means of said distributor plate, and through one of said delivery openings of said adjustable plate means, and (3) through said one of said delivery openings of said stationary plate means, through said port means of said distributor plate, and through one of said return openings of said adjustable plate means, whereby delivery of fluid to the output conduit connected to said one of said delivery openings of said adjustable plate means begins at a predetermined angular position of said rotor and terminates at varying angular positions dependent upon adjustment of said adjustable plate means.

25. Apparatus according to claim 18 wherein said openings in said plate means are angularly spaced around said axis so as to provide fluid flow from said outlet port of said rotor in the following sequence: (1) through one of said return openings of said stationary plate means, (2) through one of said delivery openings of said stationary plate means, through said port means of said distributor plate, and through one of said return openings of said adjustable plate means, and (3) through said one of said delivery openings of said stationary plate means, through said port means of said distributor plate, and through one of said delivery openings of said adjustable plate means, whereby delivery of fluid to the output conduit connected to said one of said delivery openings of said adjustable plate means begins at varying angular positions of said rotor dependent upon adjustment of said adjustable plate means and terminates at a predetermined angular position of said rotor.

26. Apparatus according to claim 18 having a passage in said distributor plate connecting said port means of said distributor plate to said fluid source, said passage containing a spring-loaded check valve.

27. Apparatus according to claim 26 wherein the portion of said passage containing said check valve extends at least partially radially in said rotor, whereby centrifugal force affects the opening pressure of said check valve.

28. Apparatus according to claim 26 wherein the portion of said passage containing said check valve extends at least partially non-radially in said rotor, whereby angular acceleration of said rotor affects the opening pressure of said check valve.

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