US2934053A - Fuel injection system - Google Patents

Description

April 26, 1960 A. HossAcK 2,934,053

FUEL INJECTION SYS'IEM Filed Sept. 30, 1958 Y 3 Sheets-Sheet l y. I3 I,lfi l 1; n 5**- l5 l le Il q; 7 114i n Z4 25. /25 1m 54 al Q $1 L i 55 so i@ i e INVENTOR. LExA/of fossAc/ BY www April 26, 1960 A. HossAcK FUEL INJECTION SYSTEM Filed sept.A so, 195e 3 Sheets-Sheet 2 April 26, 1960 A. HossAcK FUEL INJECTION SYSTEM 3 Sheets-Sheet 3 Filed Sept. 30, 1958 VlOFa'l NVENTOR. ,44pm/voe? 'l1/Ossau( Arroe/veys limited States atetit FUEL INJECTION SYSTEM Alexander Hossack, Pleasantville, N.Y., assignor to Simmonds Aerocessories, Inc., Tarrytown, NX., a corporation of New York Application September 30, 1958, Serial No. 764,421

9 Claims. (Cl. 123-139) This invention relates to a fuel injection system including a hydraulic pumping device and more particularly to an improved fuel injection system wherein the pumping device is isolated from pressure wave propagation.

Fuel injection systems have been employed with internal combustion engines for a number of years to replace carburetors. For certain applications a fuel injection system provides a much better means of delivering an v accurate charge of fuel to the cylinders of an internal combustion engine than does a carburetor type of device because o'f the ability to control more accurately the charge of fuel which is mixed with the air to make up the charge to the cylinders.

Fuel injection systems are generally designed to be controlled by three variables: speed of the engine, intake manifold pressure and intake manifold temperature. The hydraulic pumping device of a fuel injection system must automatically'produce a metered quantity of fuel which is compensated over the rangeof variations in speed, pressure and temperature which is to' be expected with the type of internal combustion engine employed.

One of the most serious difficulties with known fuel injection systems has been that pressure waves caused by the intermittent type of ow in the outlet lines and nozzles required have interfered with the operation of the pumpf ing device at certain speeds to cause a loss of eiciency and output.

It is one object of this invention to provide means for isolating the hydraulic pump component of a fuel injection system from pressures created by outlet lines and nozzle configurations.

Another object is to' provide means within the hydraulic pump component of a fuel injection system which prevents the variable elements of the pump from being affected by pressure waves present in the outlet lines.

A still further object is to prevent the pump chambers from communicating with the outlet lines when the pressure created in each pump chamber is less than the maximum pressure which is present in the outlet lines.

Further objects and features will become apparent in the full description of the pumping device taken in co'njunction with the drawings in which:

Fig. 1 is an elevation view in cross section showing a fuel injection pump;

Fig. 2 is an enlarged cross-sectional view of the pump shown in Fig. 1 on the line corresponding to 2 2;

Fig. 3 is a schematic diagram illustrating the relationship between a distributor valve port and the o'rice of a pump chamber of the pump illustrated in Fig. l just prior to and at the end of injection;

Fig. 4 is a schematic diagram illustrating the cycle of each chamber of the pump illustrated in Fig. 1;

Fig. 5 is a graph illustrating speed efect; and

Fig. 6 is a graph illustrating the variations in pressure created by a single pump chamber and the pressure wave induced in the system by the intermittent type flow.

Referring now to Fig. l. The fuel injection pump 1 that is illustrated is provided with a housing 2, first and second operating chambers 3, 4 and a capsule pressure chamber 5. A main drive shaft 6 passes through each o'f the chambers 3, 4 and provides the rotational force which operates the various components of the fuel injection pump 1. High pressure oil is forced into the chamber 3 through an oil supply connection 7. The oil in the chamber 3 passes through the hollow portion 8 of the shaft 6 and is admitted to the chamber 4 by means of a valve 9. The opening between the valve 9 and the upper end 10 of the shaft 6 is controlled by the connecting link 11 which is in turn operated by a capsule assembly 12 within the chamber 5.

The capsule assembly 12 will position the valve 9 in such a manner as to compensate for variations in manifold temperature and manifold air pressure. A manifold pressure connection 13 co'nveys air to the interior of the chamber 5 thus creating air pressure around the capsule assembly 12. T'ne interior of the capsule assembly 12 communicates directly with a capillary tube 14 which in turn is connected to a temperature bulb 15 which senses the temperature of the air in the manifold of the engine. Thus the capsule assembly 12 will be contracted or extended an amount corresponding to a balance between the effects of the manifold pressure and the manifold temperature. Since the connecting link 11 is rigidly attached to the capsule assembly 12 on one end and the valve 9 on the other the contraction and expansion of the capsule assembly 12 will control the opening between the two members and hence control the flow o'f oil from chamber 3 to chamber 4. An oil drain connection 16 restricts the flow of oil from the chamber 4 thus maintaining a static pressure within the chamber 4 which will be less than the pressure of the oil within the chamber 3 and will vary depending upon the opening between the valve 9 and the upper end 10 of the shaft 6.

A skew shaft 17 which is slidably mounted o'n the shaft 6 forms the lower end of chamber 4. The skew shaft 17 has a forked end portion 18 which is driven by the lobe portion 19 of the shaft 6. The skew shaft 17 is free to move in a vertical direction under the inuence of the pressure of the oil in chamber 3 and the pressure of the oil within the chamber 4 and the fo'rce of the tuning springs 20. Since there isa pressure closing the opening between the valve 9 and the upper end 10 of the shaft 6, the tuning springs 20 are provided which creates a force to cause a balance of total force operating on the skew shaft- 17. Thus the vertical position of the skew shaft 17 is determined by the position of the capsule as sembly 12, which is in turn, as mentioned above controlled by the manifold pressure and manifold temperature.

Slidably mounted on the skewed portion 21 of the skew shaft 17 is a ring 22. A wobble plate 23 is slidably mounted around the periphery of the ring 22. The wobble plate 23 is the means used to provide a reciprocating motion to the plunger elements 24 in the pump chambers 25. The to'tal amount of displacement during the reciprocatory motion of the plungers 24 will be determined by the angular relationship between the longitudinal axis of the wobble plate 23 and the shaft 6. Thus when the skew shaft 17 is forced upward by an increase in pressure in chamber 3 over the combined forces ex erted by the pressure in the chamber 4 and the tuning springs 26 the angle between the longitudinal axis of the shaft 6 and aline drawn along the longitudinal axis of the ring 22 will decrease and will cause a corresponding decrease in the total displacement of the plungers 24 in the chambers 25.

The skew shaft 17 is rotated at the same speed as the shaft 6 which is in turn driven by a direct connection to the engine shaft. Thus it can be seen that the plungers 24 will operate in a sequence corresponding to the rotation of the shaft 6. The shaft 6 has an eccentric portion spaanse one of the outlet passages 30 formed in the end cap 31 of the pump 1.

Fuel is supplied to the pump chambers 25 through a fuel supply inlet 32 and during the eccentric movement of the distributor valve 27 is permitted to occupy an annular chamber 33 which communicates with the pump chambers 25. Seal 34 prevents the entry of the oil into the chamber 33.

; Outlet lines 35 are connected to theV outlet passages 36 to the end cap 31. The outlet lines 35 convey the fuel from the outlet passages 30 to nozzles (not shown) which are placed in the intake portions of each of the cylinders of an internal combustion engine.

The numberA of pump chambers 25 in the pump 1, illustrated in Fig. 1 may be any convenient number. However, it has been found convenient to have each pump chamber 25 provide the charge of fuel for two cylinders. Thus for a six cylinder engine there would be three pump chambers 25, an eight cylinder engine would require four pump chambers anda twelve cylinder would require six pump chambers. With this arrangement it is necessaryl to provide suicient discharge ports 28 in the istributor valve 27 and to time the rotation of the discharge valve 27 by a gearing arrangement so as to provide for the discharge of fuelfrom the pump chambers 25 to the outlet passage 39 and hence to outlet lines 35 in the timed sequence which corresponds to the firing order of the cylinders supplied by each outlet line 35.

Thepump 1, is one which is designed to operate in a manner such that with one full revolution of the shaft 6, each plunger 24 completes one full excursion of the pump chamber 25. In order for the fuel pumped to reach the outlet components'it is necessary that there be an open passageway between the outlet components and Y the plunger 24. As may be seen in Figs. 2 and 3 the discharge ports 28 in the distributor valve 27 and the timing of the rotation of distributor valve 27 in its eccentric path, is designed to present an opening beneath each of the pump chambers 25 so that during the period of the downward movement of the plunger 24 theV pump shaft 6 makes one-half of one revolution. Thus as indicated in Fig. 3 the relationship between discharge portsr28 and the orifice 29 was such that at 0 on Fig. V4 the injection began and at 180 the injection ended. The position of the discharge port 2S is as indicated in dotted lines in Fig. 3 at the beginning and at the en`d of the injection period.

The above vdescribed pumping system has been found to be highly satisfactory at low operating speeds. The pump provided a charge of fuel to each of the cylinders of the combustion engine to which the pump was connected which was precisely metered Vin accordance with the particular engine requirements. Thus for a given manifold temperature and pressure the volumetric ow of fuel from thepump would be indirect proportion to the outlet components.

the speed of the rotation of the lshaft 6 which is directly coupled to the crank shaft of the engine.

The line A on Fig. 5 illustrates the linear relationship between theengines speed and the fuel delivery under ideal conditions with a fixed manifold pressure and fixed manifold temperature. Y However, in practice it was found that at certain speeds, instead ofthe fuel delivery continuing to increase as engine speed increased, at point B a peak was reachedas illustrated on the line C at which the fuel delivery dropped off over the rangeras illustrated and then began to increase again. This effect rendered the pump efficiency not entirely satisfactory for engines operatingwithinthis range of speeds. Thus, a variable which could not be compensated for by changes in the various components of the pump 1 was observed which will be referred to hereafter as speed eect. The result of speed effect is a change of output characteristic in the higher engine speed ranges. It was observed by eX- perimentation that at the point B on the line C at which the speed effect became prominent the angular relationship between the wobble plate 23 and the shaft 6 was changed. Many efforts were made to determine the cause of the change of the position of wobble plate 23 at this particular point with the resultant decrease in total output of the pump 1. Y

Studies indicated that at certain points resonant effects were produced by the outlet components of the fuel injection system. It was found that attempts to reduce operating clearances within the pump components to the minimum commercially acceptable did not materially improve the results of the output of the-pump 1. It was further determined -that a differential in pressure in chambers 3 and 4 became prominent at the speed at which point B occurs on line C and that the differential in pressures was always in .the same direction. However, it was determined that the changes of these pressures were the results of the speed effect and were not the cause of the decrease in output of the pump.

Other attempts at altering various components of the pump 1 were also unsatisfactory in eliminating speed effect. lIt was not until the orifice 29 was reduced in size that `any material increase in efficiency at higher engine speeds resulted. Experimentation indicated that an optimum size of the orifice 29 would result in an increase in efficiency which was sufficient to render a pump, unsatisfactory prior to the change in the size of the orifice 29, satisfactory in operation. There are several possible explanations ,for this improvement. What follows below is one of the possible explanations. v

As may be seen in Fig'. 6 there is a constant pressure within the fuel inlet side of the pump which corresponds to the pressurel created by the usual fuel pump which forces fuel into the fuel injection pump. This pressure may be at any convenient level but as illustrated in Fig. 6 is appproximately 20 p.s.i. Thus there is always a constant force against the lower end of the plunger 24 equal to the'pressure within the inlet system of 20 p.s.i. and the force of the plunger springs 36 (substantially constant) which must beV overcome `by the force nof theV plunger in v,order toproducel flow from the pump 1 into The line D on'Fig. 6 illustrates the sinusoidal characteristic in the pressure created by a plunger 24 during its dovrnwardl movement in its chamber 25. The peak pressure created theoretically must correspond to the crack pressure of the nozzles, which is shown as 120 p.s.i. At the beginning and at the end of each compression phase lthe force created by the downward movement of the plunger'corresponds to the pressure in the outlet side of the pump or 20 p.s.i. and the force of the springs 36. As may beseen from the line D fora substantial period of time the pressure within the chamber is of a very low order, roughlyV corresponding to the inlet pressure of the fuel.

It was theorized thatconditions of resonance occurred at which the forces created by pressure waves in the outlet Y side of the pump varied theV force exerted against the plunger particularly at the end of injection. Due to the pressure balance type ofgsystem within the pump theseV forces were found to cause V,the change in the angular relationshipbetween the wobble plate 23 with'the shaft counteracting forces induced by the resonant pressure waves created inthe foutlet components, it is found that these counter-acting forces may be represented by the lines E and F as indicated in Fig. 6. Thus vat low engine speeds, for example, if a resonant condition exists and a pressure wave of the type as illustrated by line E occurs, the total counteracting force induced does not exceed that of the inlet pressure and therefore has no effect on the operation of the pump. However at greater speeds the counteracting forces may assume an amplitude and period such as illustrated by line F. If the period of the counteracting force is so phased so that the peak of the force occurs at a time when the pressure created by the downward movement of the plunger is less than the total of the induced pressure and the pressure on the inlet side of the pump, a change in angular relationship of the wobble plate 23 results. Due to resonance the effect, observed on line C in Fig. 5, is an initial drop-olf toa certain point. The volumetric flow increases from this point in a substantial linear relationship with the increase in speed of the engine. This may be explained by assuming that while conditions of resonance occur at higher engine speeds the outlet lines themselves may absorb a certain amount of the energy and prevent its transmission back to the pump plungers. Because of these considerations it was necessary to provide some means to alter the pump in a way which would take into account all of the possible explanations for speed effect.

Thus the problem is one of isolating the plunger 24 from the outlet side of the pump at the end of injection and from the counteracting force induced by the resonant condition of the pressure waves in the outlet components.

It was determined that a solution to the problem could be obtained by changing the time relationship for the injection cycle of the pumps so that a passageway between the orifice 29 and the outlet passages 30 was open over a substantially shorter period of time than had been the condition before.

Prior to the change in size, the orifice 29 had had a diameter of .203" which as indicated as on dotted lines in Fig. 3 as 29'. That sized orifice 29 resulted in an injection cycle over one-half of one revolution of the shaft 6 as indicated in Fig. 4. In order to reduce the time during which an open passageway between the outlet openings and the pump plunger exists, the diameter of the orice was reduced to .125, which proved to be more than was necessary.

Further experimentation proved that a diameter of .162" was the optimum orice size for the particular type of pump being tested. Tests made with all other components of the pump remaining exactly the same and with only the diameter of the orifice changed from .203" to .162 resulted in a substantial improvement in eiciency. The line G on Fig. may be compared with the line C which illustrates the improvement obtained by the change in orifice diameter.

The effect of changing the diameter of the outlet orifice 29 is such as to delay the beginning and end of the injection cycle to a point as illustrated in Fig. 4 as 10.9. The reduction in the diameter of the outlet orifice 29 was accomplished without changing any of the other components or the timing sequence of the other operating components of the pump 1. Thus at the beginning and end of the downward movement of the plunger 24 there is no passageway for the fuel which is being compressed by the plunger 24. A pressure builds up in accordance with the curve D as shown on Fig. 5 and it is not until the point of the rotation of the shaft.corre sponds to 10.9 that iiow can occur between the pump chamber 25 and the discharge passages 30 in the end cap 31. In addition as may be seen in Figs. 3 and 4 the reduction of diameter of the outlet orifice 29 results in the ending of the injection cycle at 169.1"` of rotation, or 10.9 prior to the time the plunger 24 has reached the lower end of its stroke. Thus at the beginning and at the end of the plunger stroke for a period of time during which the shaft rotates through an angle of 10.9 in each instance, it is impossible for any ow to occur 6 out of the pump chamber and thus it is impossible for the counteracting forces present on the outlet side of the pump to affect the operation of the pump.

Since there is no passageway by which the fuel may escape from the chambers 25 at the beginning and end of the plungers compression strokes, the effect is that the pressure of the fuel in the chamber is built up during those brief intervals. This higher pressure may cause some fuel to escape across the distributor valve 27 to the inlet side of the pump. However, it is more probable that the plungers 24 cannot move appreciably during the brief intervals in which there is no passageway. If the plungers cannot or do not move, the pressure within the chambers 25 is still raised as desired by the action of the wobble plate 23. But the wobble plate 23 momentarily may not be able to force the downward movement of the plungers 24. The result is that the clearances between the wobble plate 23 and its associated elements, which are known to exist, are taken up. This would cause a slight change in the angular relationship of the wobble plate 23 and the skew shaft 17. This slightly different angular position may explain the reason that line G of Fig. 5 is slightly displaced from the ideal line A. It is also possible that a combination of some fuel leakage and taking up of available clearances occurs during the brief intervals the plunger is closed off. However, if the delay in the opening of the passageway and its early closing permits the pressure in the chamber 25 to build up too much there may be either excessive leaking of fuel to the inlet side of the pump or impressing of the plungers 24 into wobble plate 23. Therefore the intervals during which the plungers are isolated must not be greater than necessary to permit a build up of pressure sufiicient to ensure that the pressure of the fuel in the chambers will exceed the total of reactance forces on the plungers 24 when the passageway to the outlets does occur.

It will be noted from Fig. 6 that the effect of this change of the injection period is to cut olf the sinusoidal wave illustrated by line D at a particular point which corresponds to the 10.9 change in the period of the injection'cycle. This naturally results in a loss of sorne fuel output. However, the measure of the loss is something in the order of 2% reduction in volume of fuel per injection. This very slight reduction in pump output to achieve the very material elimination of the speed effect is considered to be well worth the small loss in pump output.

Thus it will be seen that the effect of changing the diameter of the orifice 29 is to cause a slight reduction in the duration of the injection period. This reduction, as illustrated by Fig. 4, in the period of injection is from 180 of the shaft rotation to 15 82 of shaft rotation and results in a 12.1% reduction in the duration time of the.

injection. While the above figures are given with reference to a particular type of pump it will be appreciated that the invention herein is not confined to the particu lar pump which has been illustrated and described. The important factors are that there must be an isolation of the pump plungers from the outlet side of the pump to ence to a particular type of pump it will be appreciated'` that the invention herein is not limited to that particular disclosure but is only to be limited by the scope and ex,-`

tent of the appended claims.

I claim: Y

1. Hydraulic pumping meansfor delivering a metered quantity, of fluid to` each of a plurality of outlet passages comprising a pluralityof pump chambers each having a plunger'therein; means for imparting a reciprocatory motion to said plungers, outlet means for passing fluid under pressure from said pump chambers to said outlet passages and means for isolating the pump chambers from said outlet passages for a short period of Vtime after the beginning of and prior to the end of said plungers compression stroke. Y

2. Hydraulic pumping means for delivering a metered quantity of tiuid to each of a plurality of outlet passages comprising a plurality of pump chambers, plungers for said chambers, means forimparting a reciprocatory motion to said plungers, orices forming a passageway Y from said chambers tos'aid outlet passages, and valve means for opening said orifices for an interval of time which is less than'the interval of time of the compression stroke of said plungers.

3. Hydraulic pumping means for delivering a metered quantity of liuid to each of a plurality of outlet passages comprising a plurality of pump chambers, plungers for said chambers, means for imparting a reciprocatory motion to said plungers, orifices forming a passageway from said chambers to said outlet passages, and valve means for opening said orifices, said orifices and corresponding valves being of such size that the orifices are opened after the start of the compression stroke of the corresponding plunger and are closed prior to the end of such stroke.

4. Hydraulic pumping means for delivering a metered quantity of fluid to each of a plurality of outlet passages comprising a plurality of pump chambers each having a plunger therein, orifice means formed in each of said chambers, means for imparting a reciprocatory motion to said plungers, valve means having discharge ports formed therein positioned between said orifice means and said outlet passages. means for rotating said valve means in a timed relationship with the reciprocatory motion of said plungersV adapted to position one of saidY ports beneath the orifice of each of said chambers and to provide a passageway from said chamber to an outlet passage for` a period of time which is less than the time required for the plunger to complete the compression stroke of its operation.

A 5. ln a fuel injection system for an internal combustion engine having a reciprocatory pump device for de-` livering a metered quantity of fuel to each of the cylinders of said engine, said pumping device comprising a plurality of stationary pump chambers, a reciprocatory plunger and an orifice for each of said chambers, rotatable pump shaft means driven by the crankshaft of said engine at a rotational speed corresponding to the rotational speed of said engine, means driven by said pump shaft means for effecting the reciprocatory movement ofsaid plungers in a timed sequence, rotatable valve means driven by said pumping shaft means, stationary distributor means having a plurality of outlet passages formed therein for the delivery of fuel to outlet lines connected to said engine cylinders, said valve means being v disposed.Y between said outlet passage and the orifice of t said pump chambers and being rotated at a speed capable of permitting the passage of fuel from each of said pump chambers to each of said outlet passages in a timed sequence and over a period of time less than the period of time required for the pressure stroke of said plungers. 6. In a fuel injection system for an internal combustion engine having a reciprocatory pump device for delivering a metered quantity of fuel to each of the cylinders of said engine, said pumping device comprising a plurality of stationary pump chambers, a reciprocatory plunger and an orifice for each of said chambers, rotatable pump shaft means driven by the crankshaft of said engine at a rotational speed corresponding to the rotational speed'of said engine, rneans'driven by said pump shaft means for eecting the recprocatory movementof said plungers ina timed sequence, rotatableV valve means driven by said pumping shaft means, stationary distributor means having a plurality of outlet passages formed therein for the delivery of fuel to outlet lines connected to said engine cylinders, said valve meansbeing disposed between said outlet passage and the orifice of said pump chambers and being rotated at a speed capable of permitting the passage of fuel from each of said pump chambers to each of said outlet passages in a timed sequence and over a period of time less than the period of time requiredl for the pressure stroke of said plungers, said discharge passage and orifices being so sized and the rotation of said valve being so interrelated that the valve does not open the orifice until an interval of'time after the plunger has started its compression stroke and that the valve closes the discharge orifice an interval of time prior to the completion of such plunger stroke.

7. In a fuel injection system for an internal combustion engine having a reciprocatorypump device for delivering a metered quantity of fuel to each of the cylinders of said engine, said pumping device comprising a plurality of stationary pump chambers, a reciprocatory plunger and an orice for each of said chambers, rotatable pump shaft means driven by the crankshaft of said engine at a rotational speed corresponding to the rotational speed of said engine, means driven by said pump shaft means for effecting the reciprocate-ry movement of said plungers in a timed sequence, rotatable valve means driven by said pumping shaft means, stationary distributor means having a plurality of outlet passages formed therein for the delivery of fuel to outlet lines connected to said engine cylinders, said valve means being disposed between said outlet passage and the orifice of said pump chambers and being rotated at a speed capable of permitting the passage of fuel from each of said pump chambers to each of said outlet passages in a timed sequence and over a period of time less than the period of time required for the pressure stroke of said plungers, said discharge passage and Vorifices being so sized and the rotation of said valve being so interrelated as to prevent the discharge of fluid from said chambers to said outlet passages when the pressure of the fluid on the discharge side of said oritice'would affect the operation of said plunger.

8. In a fuel injection system for an internal combustion engine having a reciprocatory pumping device for delivering a metered quantity of fuel to a plurality of fuel lines, said fuel lines being connected to a correspending number of injection nozzles inthe fue-l ports of said engine, said pumping device comprising a plurality of stationary pump chambers, a reciprocatory plunger and orifice for each of said chambers,'rotatable pump shaft means driven by the main shaft of said engine at a rotational speed corresponding to the rotational speed of said engine, variable means driven by said pump shaft means for effecting a reciprocatory movement of said plungers, and the displacement of said plungers, each of said plungers completing one full cycle of operation for each complete rotation of said pump shaft, rotatable valve means driven by said pump shaft and having a plurality of ports formed therein, stationary distributor means having a plurality of outlet passages formed therein for the delivery of fuel to said outlet lines andV having means for introducing fuel from the fuel supply of said engine to said pumping means, said valve means being positioned between said outlet passages and said orifices and being rotated at a speed wherebyV the passage of fuel from said pumping chambers to said outlet passagesris permitted for a period of time which is less than the period of time required` for Vthe pressure stroke of each of said plungers, said ports and said oriiices being so sized and the rotation of said valve being so timed as to prevent the discharge of fuel from saidl chambers to* said outlerpassages when the pressor' fuel iines, said fuel lines being connected to a correspending number of injection nozzles in the fuel ports of said engine, said pumping device comprising a plurality of stationary pump chambers, a reciprocatory plunger and orice for each of said chambers, rotatable pump shaft means driven by the main shaft of said engine at a rotational speed corresponding to the rotational speed of said engine, skew shaft means connected to said pump shaft means naving an olset operation, plate means mounted on said skew shaft means for eecting reciproeatory movement to said plungers, means controlled by the manifold air pressure and temperature for varying the position of said plate means with respect to said plungers whereby to vary the displacement of said 20 plungers, rotatable valve means driven by said pump shaft and having a plurality of ports formed therein, stationary distributor means having a plurality of outlet passages formed therein for the delivery of fuel to said outlet lines and having means for introducing fuel from the fuel supply of said engine to said pumping means, said valve means being positioned between said outlet passages and said orifices and being rotated at a speed whereby the passage of fuel from said pumping chambers to said outlet passages is permitted for a period of time which is less than the period of time required for the pressure stroke of each of said plungers, said ports and said orifices being so sized and the rotation of said valve being so timed as to prevent the discharge of fuel from said chambers to said outlet passages when the pressure of the fuel on the outlet passage side of said orice would affect tbe operation of said means for controlling the displacement of said plungers.

eferenees Cited in the file of this patent UNITED STATES PATENTS 2,697,401 Barberena Dec. 2l, 1954 2,764,964 Meyer Oct. 2, 1956 FOREIGN PATENTS 405,803 France Nov. 24, 1909

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