US3505984A

US3505984A - Fuel injection systems for internal combustion engines

Info

Publication number: US3505984A
Application number: US734442A
Authority: US
Inventors: Wilfred Percival Mansfield
Original assignee: British Internal Combustion Engine Research Institute
Current assignee: British Internal Combustion Engine Research Institute
Priority date: 1967-06-05
Filing date: 1968-06-04
Publication date: 1970-04-14
Anticipated expiration: 1987-04-14
Also published as: FR1566995A; GB1232973A; DE1751474A1

Description

FUEL INJECTION SYSTEMS FOR INTERNAL COMBUSTION ENGINES Filed June 4, 1968 April 14, 1970 w. P. MANSFIELD 5 Sheets-Sheet 1 wrz/vme WILFRED F! MANSFIELD. *%*2"* 4770mm? APril w. P. MANSFIELD 3,505,984

FUEL INJECTION SYSTEMS FOR INTERNAL COMBUSTION ENGINES Filed June 4, 1968 v 5 Sheets-Shet 2 FIG.3.

FUEL INJECTION SYSTEMS FOR INTERNAL COMBUSTION ENGINES Filed June 4. 1968 April 14, 1970- w. P. MANSFIELD 5 Sheets-Sheet 5 VVV////////////////////// April 14, 1970 w. P. MANSFIELD FUEL INJECTION SYSTEMS FOR INTERNAL COMBUSTION ENGINES 5 Sheets-Sheet 4.

Filed June 4, 1968 April 14, 1970 w. P. MANSFlELD' 3,505,984

FUEL INJECTION SYSTEMS FOR INTERNAL COMBUSTION ENGINES Filed June 4, 1968 I 5 Sheets-:Sheet 5 United States Patent Office 3,505,984 Patented Apr. 14, 1970 3,505,984 FUEL INJECTION SYSTEMS FOR INTERNAL COMBUSTION ENGINES Wilfred Percival Mansfield, Slough, England, assignor to The British Internal Combustion Engine Research Institute Limited, Slough, England, a British company Filed June 4, 1968, Ser. No. 734,442 Claims priority, application Great Britain, June 5, 1967, 25,879/67 Int. Cl. F02m 41/00 US. Cl. 123-139 11 Claims ABSTRACT OF THE DISCLOSURE A fuel injection system for a multi-cylinder internal combustion engine in which the injector nozzles in the cylinders are fed sequentially by fuel under pressure to open the nozzle valve in each, by way of a plunger operating in a bore positioned coaxially within a housing rotatable in a casing, cooperating ports and passages being provided around the axis of the rotatable casing and housing controlling the passage of fuel to the nozzles, the plunger being operated by liquid under pressure from a storage accumulator by means of a servopiston also co-axial with the bore in the housing, with said servo-piston being of greater diameter than that of the bore for the plunger. The pressure fluid is spilled to a source of lower pressure at the completion of injection into each cylinder.

This invention relates to fuel injection systems for internal combustion engines.

The object of the invention is to eliminate the complicated and expensive mechanically operated pumping arrangements of the known distributor type fuel injection-pump and to substitute a hydraulically operated pump system. Further objects of the invention are to employ the known type of hydraulic pressure intensifier in a novel manner in a distributor type pump and to make advantageous use of the methods and means of controlling the fuel injection operation which are made possible or facilitated by this system. The application of the known distributor type pump has been limited to relatively small engines and another object of the invention is to remove this limitation so that the distributor system can be applied to medium and large sizes of engines. A still further object is to eliminate the requirement in hydraulically operated systems of providing a separate injection pump for each cylinder of the engine.

In the following description these objects are achieved by employing an external pump which raises the pressure of the servo-liquid to a fraction only of the required injection pressure, and this pressure liquid is stored in an accumulator and then delivered under precise control to actuate a servo piston which has a correspondingly greater area than the injection pump plunger operated thereby whereby the fuel to be injected is raised to the required injection pressure.

This arrangement has been selected as the most practical system but it will be understood by those skilled in the art that the servo-liquid could be raised to therequired injection pressure by the external pump so that the servo piston of larger area than the injection pump plunger would be unnecessary if the supply of servo liquid from the external pump could be timed to correspond with the requirements for operating the injection pump plunger in timed sequence with the cycle of operations of the engine so that the accumulator would be unnecessary.

The invention consists in a fuel injection system for an internal combustion engine having at least two cylinders comprising a source of liquid under pressure, a liquid pressure operated fuel injection pump plunger, a housing having a bore in which the plunger works, a source of fuel and channel means connecting said source of fuel with said bore, wherein the improvement comprises providing the casing with a cylindrical bore within which the housing is rotatable and mounting the plunger and bore oo-axially within the housing, and providing cooperating ports and passages around the axis of the ro tatable casing and housing for the delivery of fuel under pressure sequentially to the front end of an injector nozzle in each cylinder of the engine to open the nozzle valve and cause injection of fuel into each cylinder, a storage accumulator means connected to said source of liquid under pressure, a liquid pressure operated servopiston working in a cylinder of larger diameter than said bore and actuating said plunger, and connected to the storage accumulator means, and spilling means for spilling the pressure fuel from the passages to a zone of lower pressure at the completion of injection into each cylinder.

The invention further consists in a fuel injection system as set forth in the preceding paragraph in which a plunger executes one pumping stroke per cylinder per cycle of said engine.

The invention still further consists in a fuel injection system as set forth in the preceding paragraph in which after the end of each injection, the liquid under pressure which operates a fuel injection pump is discharged to a zone of lower pressure under the control of passages in said housing and casing, which communicate with each other in timed sequence with the cycle of operation of the engine, and leads to an adjustable throttle.

The invention still further consists in a fuel injection system as set forth above in which the liquid under pressure that operates the fuel injection pump is fuel.

The invention still further consists in a fuel injection system as set forth above in which additional passages in said housing are placed in communication with additional passages in said casing through which, in timed sequence with the cycle of operations with the engine, high pressure fuel from said bore is led to a pipe communicating with each injector nozzle mounted in a cylinder of the engine, and is delivered to the rear end of said nozzle to co-operate in closing the nozzle valve and terminating the injection of fuel into said cylinder, and in maintaining said nozzle valve closed until injection is required on the next cycle, and in which means are provided for spilling said high pressure fuel from said passages to a zone of lower pressure prior to injection on the next cycle.

The invention still further consists in a fuel injection system as set forth above in which means are provided for automatically varying the timing of injection in accordance with the engine speed, with said means comprising a plunger slidable in a bore provided with a groove with increasing cross-sectional area which constitutes an escape path for a flow of liquid under pressure, a quantity of which is dependent on engine speed, and in which the force exerted by the flow of liquid under pressure is applied to one end of said plunger, and is resisted by the force of a spring applied to the other end of the plunger, so that, as the speed increases and the flow of liquid pressure increases, the plunger is moved axially against the force exerted by the spring until the crosssectional area of the groove is increased to a value sufficient to discharge the flow of liquid under pressure, and said axial movement in conjunction with helical grooves in the driving or driven shaft causing angular variation of the driven shaft relative to the driving shaft.

The accompanying drawings show, by way of example only, embodiments of the invention, in which:

FIGURE 1 is a longitudinal section through an injection pump constructed in accordance with the invention;

FIGURE 2 is a cross-section through the pump of FIG- URE 1 in line with the supply and discharge ports of the servo-cylinder, and including an adjustable throttle valve;

FIGURE 3 is a section on the line II-II of FIGURE 2, showing the adjustable throttle valve;

FIGURE 4 is a section through an improved injector constructed in accordance with the invention;

FIGURE 5 is a section through an alternative injector constructed in accordance with the invention;

FIGURE 6 is a cross-section through the pump in line with the speed control valve of an all-speed and torque control governor;

FIGURE 7 is a longitudinal section on the line III-III of FIGURE 6, showing the governor system;

FIGURE 8 is a part longitudinal section showing the governor system positioned to control the quantity of fuel supplied to the injection pump during the pump filling period;

FIGURE 9 is an end elevation of the pump for a six cylinder engine viewed from the driven end.

By way of example, FIGURE 1 shows one arrangement of a fuel injection pump constructed in accordance with the present invention. A cylindrical housing 1 is mounted for rotation within a non-rotating casing 2 so that it operates as a rotary valve and is driven from the engine. Axially arranged in the housing 1 are a servo-cylinder 3 and a servo-piston 4 which actuates an injection pump plunger 5 in its bore 6.

A servo-pump and accumulator of known type may be provided. The servo-pump normally comprises a single piston operated by a single eccentric or cam and is driven at engine speed. The size of the accumulator may be reduced by arranging the servo-pump piston to be driven by a multi-lift cam having a number of cam lifts. For example, there may be the same number of cam-lifts as there are cylinders in the engine, or any arrangement of cams in conjunction with their driven speed, which provides the same number of discharges of servo-liquid from the pump to the accumulator as the number of discharges from the accumulator to the servo-cylinder. Each dis charge from the servo-pump into the accumulator may occur at substantially the same time as each discharge from the accumulator into the servo-cylinder. In a further arrangement, the servo-pump may be provided with one or more pairs of diametrically opposed pistons, so that the forces on the driving shaft are substantially balanced, and which are operated by one or more eccentrics or cams driven from the engine at the speed ratio necessary to give the timing of the discharges from the servo-pump as set out above. It will be understood that this timing of the discharges is advantageous but is not essential for the satisfactory operation of the injection system.

The servo-liquid may be fuel but if it is desired to operate the engine on a heavy fuel which is unsuitable as a servo-liquid this may be done by providing separate circuits for the heavy fuel and the servo-liquid. The heavy fuel may be supplied at a suitable pressure by a feed pump to the inlet valve of the injection pump chamber so that the pipes to each injector are charged with heavy fuel, while a light fuel, or special hydraulic fluid, may be used in that part of the system which operates the injection pump plunger.

Whichever arrangement of servo-pump is used it must supply servo-liquid under pressure to the servo-cylinder 3 through a connection 7a and a supply port 7. The housing 1 is provided with equally spaced ports 8, one for each cylinder of the engine. Servo-liquid under pressure is supplied through the port 8 to the cylinder 3 to force the piston 4 to the right and actuate the plunger 5. In a twostroke cycle engine, the housing 1 is rotated at engine speed and in a four-stroke cycle engine at half-engine speed. In this way, the pump plunger 5 is actuated once per cycle for each cylinder and in timed sequence with the cycle of operations of the engine. More than one supply port 7 may be provided if desired.

Fuel is supplied under pressure, which is preferably servo-pressure, to the plunger bore 6 to the right-hand side of plunger 5, through a passage 9 and annular groove 9a in the non-rotating casing 2, radial passages 10 in the housing 1, and a spring loaded feed valve 11. When the plunger 5 is actuated, as described above, and moves to the right, the fuel is discharged at high pressure through a single radial passage 12 which communicates in turn with each of a number of outlets 13, and there is one outlet for each cylinder of the engine. From the outlet 13 a pipe 13a conveys the high pressure fuel to an injector at least one of which is mounted in each cylinder of the engine. The rate of discharge of fuel to the passage 12 throughout the injection period, can be controlled to a certain extent by choice of the geometry of the supply port 7 and ports 8.

Injection is terminated when movement of the plunger 5 to the right closes the passage 12 (the position shown) and opens communication between the passage 12 and an annular groove 14 in the plunger 5. The groove 14 communicates through passage 15 with another groove 16 in the plunger 5, and at this time the groove 16 is placed in communication with a radial passage 17 which communicates with a groove 18 around the housing 1. In this way fuel from the injector and its associated pipe is spilled to a zone of low pressure, for example, the sump of the servo-pump or other reservoir through

passages

19 and 20 in the casing 2. To control the rate at which the fuel is spilled, a spring-loaded valve 21 may be positioned between the

passages

19 and 20. The valve 21 may be arranged to control both the pressure at which it opens and the maximum area of opening. In FIGURE 1 the pressure exerted by the spring upon the ball and the lift of the ball are shown to be independently adjustable but this is for experimental purposes and when the required values have been ascertained, a simple arrangement giving fixed values of these features can be employed on all further pumps for a given application.

As soon as the pressure at the end of the bore 6 falls below the servo-pressure, in the manner described later, the feed valve 11 opens so that further fuel is supplied under servo-pressure to the plunger bore 6, as described above, and moves the plunger 5 and servo-piston 4 to the left thus displacing servo-fluid from the cylinder 3. To permit this displacement, the casing 2 is provided with a discharge port 22 shown in FIGURE 2, which is arranged to communicate with the ports 8 in the housing 1 at the required times in the cycle of operations of the engine. The liquid discharged from the port 22 is returned to the servo-pump sump or other reservoir through a simple adjustable throttle such as is shown in FIGURE 2. The throttle comprises an H-valve 23 operating in a bore 24 and arranged to open and close the port 22. Movement of the H-valve 23 downwards in this figure progressively closes the port 22 and has the effect of reducing the volume of fuel injected into the cylinder of the engine as described below. The movement is resisted by a spring 25 and is caused by a projecting segmental portion 26 of a control spindle 27, shown in FIGURE 3, which is turned by a control lever 28. The portion 26 limits the spring assisted movement of the H-valve 23 upwards. Movement downards is limited by an idling speed adjusting screw 29 loaded by a spring 30 and mounted in a screwed sleeve 31 which is slidable in a plug 32 and which is provided with a sealing cap 33. Additional force applied to the lever 28 moves the H-valve further downwards against the resistance of the two

springs

25 and 30 thus closing the port 22 completely and shutting down the engine. In addition to being controlled manually, the throttle may be controlled by an engine governor of the maximum speed or all-speed types, or by a maximum area limiting device depending on the type of service for which an engine is required. Variation of the area of the passageway through the throttle determines the volume of liquid that is discharged from the cylinder 3 and thus the length of the return movement to the left in FIGURE 1 of the piston 4. The movement of the plunger is the same and thus determines the volume of fuel that is passed into the bore 6 and is injected into a cylinder of the engine on the next injection. The arrangement ensures that the required volume of fuel is present in the bore 6 prior to one of the ports 8 coming into communication with the supply port 7 to start a new injection. The arrangements used to meter the volume of fuel and to cause the injection of fuel are the same on all injections, so each cylinder receives the same volume of fuel under the same conditions. The arrangement also ensures that the maximum return movement of the plunger 5 is completed before the passage 12 uncovers the next port in the casing 2 leading to the injector of another cylinder.

FIGURE 1 shows a preferred arrangement of the servo-system in which the servo-cylinder 3 and servopiston 4 are of larger diameter than the pump plunger 5 and its bore 6. In this way, the servo pressure required for a given injection pressure may be reduced or the injection pressure for a given servo pressure may be increased. This arrangement does not require any changes in the operation of the system as described above.

In accordance with a further embodiment of the invention, the cylindrical housing 1 is not rotated and a cylindrical rotary valve is mounted between the nonrotating housing 1 and the non-rotating casing 2. The rotary valve is provided with ports which place passages in the housing 1 as described above into and out of communication with passages in the casing 2 in timed sequence with the cycle of operations of the engine and as described above.

The servo-liquid pump and the accumulator may be built as a unit with the fuel injection pump and in any case it is desirable that the accumulator be adjacent to the fuel injection pump.

The normal types of injectors may be used but in accordance with another embodiment of the invention the injection pump described above is used in conjunction with a special design of injector. In modern engines there is often considerable difficulty in arranging suflicient space in the cylinder head to accommodate the injector. Thus injectors of small dimensions are coming into use and the present proposals facilitate the employment of small injectors. The injector shown in FIGURE 4 includes a standard nozzle 40: and cap nut 41, as it is of practical importance that standard nozzles can be used. In addition to the nozzle, the injector comprises a body 42 in which are arranged

connections

43 and 44 and

passages

45 and 46. The connection 43 and passage 45 lead to a lifting face 47 of a needle valve 48 in the normal manner. The connection 44 and passage 46 lead to the rear end of the needle valve 48. The normal push rod, spring, spring plates and spring adjusting screw have been eliminated with a consequent reduction in cost and a reduction in the mass of the moving parts. The normal leakage fuel connection is replaced by the second high pressure connection 44.

Although the injector shown in FIGURE 4 is smaller than the normal injector it is not as small as is possible with the present system and FIGURE 5 shows an altemative injector of small dimensions. This comprises a body 82 which can be manufactured cheaply from steel tubing, containing a needle valve 83, the pointed end of which seats in a nozzle 84 while the other end is located by a close-fitting bush 85 pressed into the body. When the nozzle 84 is in positiomthe end of the body 82 is rolled over. A threaded connector 86 for the pipe communicating with a space 87 at the rear end of the needle valve serves also to limit the movement of the valve, with adjustment being effected by turning the connector and then securing it in position by means of a lock-nut 88. A

banjo-type connector 89 for the pipe communicating with the front end of the valve is shrunk 'on to the body 82.

The operation of these arrangements is described in relation to FIGURE 4 and the operation in accordance with FIGURE 5 is the same. The connection 43 is connected to a pipe 13a which leads to one of the connections 13 in the casing 2, shown in FIGURE 1. When the plunger 5 is actuated and discharges fuel under high pressure, as described above, this fuel is delivered through passage 45 to the front end of the needle valve 48 thus opening the valve and injecting fuel into the cylinder in the normal manner. Supply of oil to passage 45 is terminated as described earlier, including the spilling of the fuel in the pipe 13a to a zone of low pressure.

The connection 44 is connected to a pipe 49 which leads to one of connections 50 in the casing 2. In the same manner as for connections 13, described in relation to FIGURE 1, there is a connection 50 for each cylinder of the engine and thus for each injector. The connection 50 communicates in timed sequency with the cycle of operations of the engine with a single radial passage 51 in the housing 1 which leads to the bore 6. As described earlier, in connection with FIGURE 1, movement of the plunger 5 to the right closes the passage 12 and stops delivery of fuel through the pipe 13a to the front of the nozzle 48, but the movement of plunger 5 to the right is arranged to continue so that fuel is pumped through a longitudinal passage 52 in the plunger 5 and radial connecting passage 53 to annular groove 54. When the further movement of the plunger 5 places the groove 54 in communication with the radial passage 51 fuel under high pressure is delivered through the connection 50, pipe 49, connection 44 and passage 46 to the rear end of the needle valve 48. This action is arranged to occur a little earlier than the spilling of the fuel in the pipe 13a leading to the front end of the valve 48. In this way, the pressure on the front of the valve 48 is falling back to its residual pressure determined by the spring-loaded valve 21, while the pressure on the rear end of the valve 48 is increasing rapidly. The valve 48 is forced to its seat before the pres sure at the front thereof has fallen to a low value and the fuel in front of the valve 48 is pumped out at a high pressure so that a high rate of injection is maintained until the end of injection. Shortly after this time one of the ports 8 in the housing 1 uncovers the discharge port 22 in the casing 2 so that the pressure in the bore 6 and also in the pipe 49 falls and when servo pressure is reached, the spring-loaded feed valve 11 opens and fuel at servo pressure from passage 9 maintains this pressure while pushing the plunger 5 and servo-piston 4 to the left. The pressure maintained on the rear of the nozzle valve 48 ensures that the nozzle remains closed until the next injection occurs.

Thus, all the required changes of pressure at both front and rear of the nozzle needle will have been completed by the time the plunger 5 commences its return stroke.

'Rotation of the housing 1 closes the passage 12 as soon as the pipe 13a has been spilled and a little later closes the passage 51 leading to the pipe 49. This latter action occurs shortly after the discharge port 22 is opened. The arrangement also ensures that the maximum required return movement of the plunger 5 is completed before the passages 12, 51 uncover the next pair of ports in the casing 2 leading to the injector of another cylinder.

As the servo-piston 4 approaches the end, of its stroke in moving to the right, it pumps out the leakage fuel ahead of it through passage 55, annular groove 56 and passage 57 to a spring-loaded non-return valve 58 whence it is returned to the sump of the servo-pump or other reservolr.

In an alternative arrangement, the pipes 49 leading to the rear of the nozzle valve 48 may be unloaded through an adjustable spring-loaded valve 59 which is similar to the valve 21. In this arrangement, the valve 59 communicates with annular groove 54 through passage 60, annular groove 61 and passage 62. This unloading of pipe 49 is arranged to occur very shortly after it has received the high pressure fuel. The valve 59 provides a limited escape passage for fuel during the further movement of plunger and thus forms a dash pot bringing the plunger 5 and piston 4 to rest more gradually over a greater distance. As soon as the plunger 5 comes to rest, the pipe 49 will commence to unload through the valve 59, so that more time is available for unloading and the action can be more gradual. The opening pressure of the valve 59 will be set somewhat above the servo-pressure and when the discharge port 22 is opened, the pressure in the pipe 49 and associated passages will fall to the servo-pressure. In addition, the valve 59 allows fuel to be discharged from the pipe 49 when the needle valve 48 moves to the open position. When the plunger 5 moves to the left to commence even the shortest pumping stroke, the radial passage 51 and passageway 62, 61, 60 leading to the valve 59 are placed in communication through a wide annular groove 63 so that when the next pumping stroke begins and the nozzle valve 48 moves to its open position, some fuel is displaced through the valve 59. The arrangements whereby the various actions take place more gradually than normal have a beneficial effect in reducing the amplitude of the waves created by these actions.

The torque required to drive the housing 1 is very light so the system facilitates the employment of a simple arrangement for the automatic variation of injection timing with change of engine speed. For example, a plug 64 is used to close the servo-cylinder 3 and to act as a stop for the servo-piston 4 at maximum output. This plug is secured by a threaded ring 65 and is provided with a bore 66 in which slides a close fitting plunger 67. Head 68 of the plunger 67 is provided with a headed bolt -69 which carries two pairs of rollers 70. The outer pair engage in two helical grooves 71 cut in the side of a bore in a driving shaft 72 and the inner pair engage in two axial grooves 71a cut in the side of a bore in the plug 64. The rollers 70 are retained in position by a washer and split pin at the outer end of the headed bolt 69.

Axial movements of the plunger 67 in its bore 66 produce angular displacements of the plug 64 and housing 1 relative to the driving shaft 72 and thus provide the required advance and retard of the injection timing.

To achieve the required axial movements of the plunger 67, the fuel discharged from the valve 21, and if desired also from the valve 59, is led to the right-hand end of the bore 66 via passages 20 and 73, groove 74, passage 75, groove 76 and passage 77. The frequency of the discharges of fuel from the valves 21 and 59 is proportional to the engine speed and by the use of the arrangement described, a flow of liquid under pressure, the quantity of which is dependent on engine speed, is produced and the force exerted by this flow of liquid under pressure is applied to the right-hand end of the plunger 67. An escape path for this fuel from the bore 66 to the sump or other reservoir is provided by a groove 80 formed in the plunger 67. A spring 78 located in a bore 79 in the driving shaft 72 acts on the other end of the plunger 67. The liquid under pressure entering the bore 66 moves the plunger 67 to the left against the increasing force of the spring 78 until the pressure in the bore 66 and the area of the groove 80 exposed at the open end of the bore are, in combination, sufiicient to discharge the speed-dependent flow of liquid. Any desired variation of injection timing with speed is obtained by choice of the spring 78 and selection of the variation of the cross-sectional area of the groove 80 along its length. If a further control of this timing variation is required, an adjustable or interchangeable fixed bleed 81 can be provided in the passage 20.

When the simple throttle shown in FIGURES 2 and 3 is used, the lever 28 must be operated to adjust the pump output per stroke to the amount required according to the speed and torque required of the engine. The maximum pump output obtainable is determined by the 8 maximum pump stroke, which is limited by the position of the inner face of the plug 64 in FIGURE 1. The maximum pump output is therefore approximately the same at all engine speeds.

The valve arrangements shown in FIGURES 6 and 7, provide all-speed governing of the engine to which the pump is fitted, and limit the maximum torque of the engine to any desired value at each running speed. These arrangements, like the simple throttle valve shown in FIGURES 2 and 3, control the return fiow of servo oil from the discharge port 22.

In FIGURES 6 and 7, the oil leaving the discharge port 22 passes into passage and hence, via a first throttling orifice 91 controlled by a speed control valve 92 BIIdLfl second throttling orifice 93 controlled by a torque control valve 94, to a discharge passage 95 and discharge passage outlet 96 which is connected to the reservoir.

The end of a bore 97 in which the torque control valve slides and the end of a bore 98 in which the speed control valve slides are both connected by a passage, such aspassage 99 in FIGURE 6, to a bore 100 which is connected by a passage 101 which the groove 74 already referred to in connection with FIGURE 1. By virtue of the actions already described, the pressure of the oil in this groove increases with increase of engine speed. Hence, a speed-dependent pressure acts on one end of each of the two

valves

92 and 94.

The speed control valve 92 is provided with a spring 105, the force of which, in relation to the position of the valve, is adjustable by moving a sleeve 106, e.g. by means of a lever 107 shown in section. The position of the sleeve 106 determines the engine speed at which the rising pressure at the end of the bore 98 overcomes the force of the spring and moves the valve 92 downwards, reducing the effective area of the orifice 91, thus limiting the speed.

In the case of the torque control valve 94, the force produced by the speed-dependent pressure acting on one end of the valve is opposed by the force of spring 102 housed partly within the valve and partly within a plug 103 which closes the end of the valve bore. The spring is compressed increasingly as the engine speed increases. The position of the valve, which is shown in the maximum speed position, and hence the effective fiow area of the orifice 93 are thus dependent on engine speed. By suitable choice of the shape of surface 104 of the valve, the flow areas available at the orifice 93 at various engine speeds and hence the values of maximum torque obtainable at various engine speeds may be arranged to suit any particular engine duty, within the limits of performance of the engine.

The engine may be made to idle at a selected speed by limiting the movement of the sleeve 106 or lever 107 downwards in FIGURE 7, but it has been found that steadier idling is obtained by directly limiting the movement of the speed control valve 92 downwards so that a fixed minimum effective flow area is provided at the orifice 91. A rod 108, which is adjustably supported by a bracket 109 and locked in the adjusted position by a nut 110, permits a rapid adjustment of the idling speed to a precise value by limiting the movement of the valve 92 downwards. Where ready adjustment is unnecessary, the rod 108, bracket 109 and nut 110 may be omitted, with sleeve 106 being replaced by a simple plunger, and the downwards movement of the valve 92 being limited by the inner end of a plug 111, which is suitably dimensioned.

When the speed of the engine has been reduced to idling speed, the valve 94 will have moved close to the upper end of the bore 97. To stop the engine, rod 112 is moved upwards, for example, :by means of a lever 113 shown in section, thus moving the valve 94 still further upwards so that the orifice 93 is completely closed.

Oil which leaks from the

passages

90 and 95 past the valves to the ends of the bores containing the springs is drained to the reservoir via passages 114, 115 and 116 shown in FIGURE 6.

An alternative arrangement is shown in FIGURE 8 in which instead of controlling the output of the pump by throttling the discharge flow from the servo cylinder by means of

valves

92 and 94, the valve arrangement as shown in FIGURES 6 and 7 is moved through 90 and is placed across the passage 9, through which fuel at servo-pressure passes from the accumulator to the pump inlet valve 11. In this manner the inlet flow of fuel is throttled and thus the quantity of fuel entering the pump during the filling period is controlled. The actions of the speed and

torque control valves

92 and 94 in the arrangement are the same as described above.

In a case where the servo-piston of a diameter greater than the pump plunger is used, the flow rate of the fuel to the pump chamber is a fraction only of the rate of discharge flow from the servo cylinder. The effective areas required at the orifices 91 and 93 will be correspondingly smaller and the valves themselves and the passages may be made smaller if desired.

I claim:

1. In a fuel injection system for a multi-cylinder internal combustion engine of the type comprising a fuel injector for each cylinder and a nozzle valve for each injector, a source of liquid under pressure, a casing, a housing supported within the casing, a pump having a liquid pressure operated plunger working in a bore in said housing, a supply of fuel unler pressure, channel means conveying said fuel to said pump bore, and a distribuitor controlling the passages of fuel from said bore cyclically to said injectors, the improvement comprising providing said casing with a cylindrical bore within which said housing is rotatable, the pump plunger at its bore being positioned co-axially within said housing, flow control means in the channel means for the supply of fuel to said pump bore, said housing having ports and passages, said casing having passages and said pump plunger having interconnected co-axial grooves which co-operate respectively with the ports and passages in the housing for delivery of fuel sequentially by way of the passages in the casing to said injectors, and for the sequential spilling of said fuel from said passages to a zone of lower pressure at the completion of injection into each cylinder.

2. The fuel injection system as claimed in claim 1 in which means are provided for controlling the rate of pressure fall and the value of the residual pressure during the spilling of fuel from the passages leading to the front of the injector nozzle valve.

3. The fuel injection system as claimed in claim 2 in which means are provided for controlling the rate of pressure fall and the value of residual pressure during the spilling of fuel from the passages leading to the back of the injector nozzle valve.

4. The fuel injection system as claimed in claim 1 including a cylinder, a pressure liquid operated servo-piston working in said cylinder, said cylinder being larger than said bore for actuating said pump plunger to effect injection, storage accumulator means, the cylinder of said servo-piston being connected to the source of liquid under pressure by said storage accumulator means, an adjustable throttle, and means for controlling in timed sequence with the cycle of operations of the engine the flow of liquid under pressure from said accumulator to the cylinder of said servo-piston to actuate said pump plunger to effect each injection, and after each injection, the discharge of liquid from said cylinder to a zone of low pressure through the adjustable throttle.

5. The fuel injector system as claimed in claim 4 in which the servo-piston works in a cylinder co-axial of the pump plunger and Within the housing.

6. The fuel injection system as claimed in claim 1 having further co-axial grooves and a passage in the pump plunger which grooves co-operate with ports and passages in the housing for controlled delivery of fuel under high pressure by way of passages in the casing to the rear end of each nozzle valve to assist to closing said valve and for the sequential spilling of said fuel from said passages to a zone of lower pressure at the completion of injection into each said cylinder.

7. The fuel injection system as claimed in claim 6 in which the controlled delivery of fuel to the front and rear of each nozzle valve is effected by and during the stroke of the pump plunger in such a manner that when the stroke of the plunger has completed the delivery of fuel under pressure to the front of the respective nozzle valve, continued movement of the pump plunger delivers fuel under pressure to the rear end of the nozzle valve.

8. The fuel injection system as claimed in claim 1 including means for automatically varying the timing of injection in accordance with the engine speed, said means comprising a plunger slidable in a bore and provided with a groove with increasing cross-sectional area which constitutes an escape path for a flow of liquid under pressure, the quantity of which is dependent on engine speed, and the force exerted by the flow of liquid under pressure being applied to one end of said plunger, a spring applied tothe other end of the plunger for resisting such force, so that, as the speed increases and the flow of liquid pressure increases, the plunger is moved axially against the force exerted by the spring until the cross-sectional area of the groove is increased to a value sufiicient to discharge the fiow of liquid under pressure, and said axial movement in conjunction with helical grooves in a driving or driven shaft causing angular variation of the driven shaft relative to the driving shaft.

9. The fuel injection system as claimed in claim 8 in which the adjustable throttle is an H-valve operated by a lever.

10. The fuel injection system as claimed in claim 8 in which two adjustable throttle valves are controlled by means providing all-speed governing and also torquespeed control.

11. The fuel injection system as claimed in claim 10 in which two adjustable throttle valves are controlled by all-speed governing and torque-speed control means which operate in accordance with the pressure of the liquid, with said pressure being dependent upon the speed of the engine and applied to one operative face of each said valve, relative to a spring force applied to the other operative face of said valve.

References Cited UNITED STATES PATENTS 2,291,939 8/1942 Amery 123l 39.11 2,803,234 8/1957 Mansfield l23l39 2,816,533 12/ 1957 Reggio 123l39.9 3,058,425 10/ 1962 Evans.

LAURENCE M. GOODRIDGE, Primary Examiner US. Cl. X.R. 1035