US3856438A - Fuel injection pump - Google Patents

Fuel injection pump Download PDF

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Publication number
US3856438A
US3856438A US00318297A US31829772A US3856438A US 3856438 A US3856438 A US 3856438A US 00318297 A US00318297 A US 00318297A US 31829772 A US31829772 A US 31829772A US 3856438 A US3856438 A US 3856438A
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Prior art keywords
engine
spill
fluid
fuel
plunger
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US00318297A
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English (en)
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A Simko
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Ford Motor Co
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Ford Motor Co
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Priority to US00318297A priority Critical patent/US3856438A/en
Priority to CA181,381A priority patent/CA987560A/en
Priority to GB5610673A priority patent/GB1425875A/en
Priority to JP13746573A priority patent/JPS5425569B2/ja
Priority to DE2362526A priority patent/DE2362526A1/de
Application granted granted Critical
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/22—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves
    • F04B49/225—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves with throttling valves or valves varying the pump inlet opening or the outlet opening
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D1/00—Controlling fuel-injection pumps, e.g. of high pressure injection type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M59/00—Pumps specially adapted for fuel-injection and not provided for in groups F02M39/00 -F02M57/00, e.g. rotary cylinder-block type of pumps
    • F02M59/02—Pumps specially adapted for fuel-injection and not provided for in groups F02M39/00 -F02M57/00, e.g. rotary cylinder-block type of pumps of reciprocating-piston or reciprocating-cylinder type
    • F02M59/04—Pumps specially adapted for fuel-injection and not provided for in groups F02M39/00 -F02M57/00, e.g. rotary cylinder-block type of pumps of reciprocating-piston or reciprocating-cylinder type characterised by special arrangement of cylinders with respect to piston-driving shaft, e.g. arranged parallel to that shaft or swash-plate type pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M59/00—Pumps specially adapted for fuel-injection and not provided for in groups F02M39/00 -F02M57/00, e.g. rotary cylinder-block type of pumps
    • F02M59/20—Varying fuel delivery in quantity or timing
    • F02M59/36—Varying fuel delivery in quantity or timing by variably-timed valves controlling fuel passages to pumping elements or overflow passages
    • F02M59/361—Valves being actuated mechanically
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M59/00—Pumps specially adapted for fuel-injection and not provided for in groups F02M39/00 -F02M57/00, e.g. rotary cylinder-block type of pumps
    • F02M59/20—Varying fuel delivery in quantity or timing
    • F02M59/36—Varying fuel delivery in quantity or timing by variably-timed valves controlling fuel passages to pumping elements or overflow passages
    • F02M59/361—Valves being actuated mechanically
    • F02M59/362—Rotary valves
    • F02M59/363—Rotary valves arrangements for adjusting the rotary valve
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D2700/00—Mechanical control of speed or power of a single cylinder piston engine
    • F02D2700/02—Controlling by changing the air or fuel supply
    • F02D2700/0269—Controlling by changing the air or fuel supply for air compressing engines with compression ignition
    • F02D2700/0282—Control of fuel supply
    • F02D2700/0284—Control of fuel supply by acting on the fuel pump control element
    • F02D2700/0287—Control of fuel supply by acting on the fuel pump control element depending on several parameters

Definitions

  • the cam contour 2,992,619 7/1961 Nilges 91/499 is arranged to provide for a rate of plunger displace- 3,045,604 7/1962 Hahn 74/60 ment which is increasing for a major portion of the 3,046,950 7/1962 Smith 91/491 180 angular pump shaft rotation preceding the 3,319,568 RCpkO 6t 31. plunger top dead enter position
  • the cam surface is FOREIGN PATENTS OR APPLICATIONS radial with respect to its axis of rotation and the l 914 598 8/1970 German 91/499 plungers are provided with means to engage the cam y surface along a radial line and with means to prevent the plungers from rotating about their own axes.
  • the present invention is related to the field of piston, or plunger type, fuel injection pumps for internal combustion engines operating on the Otto cycle. More particularly, the present invention is related to that portion of the above noted field in which a single actuator reciprocates a plurality of plungers. Specifically, the pres ent invention is related to that portion of the above noted field concerned with fuel injection pumps for constant air/fuel ratio operation of the associated engme.
  • such pumps provide a plurality of piston actuated pumping means arranged around a central bore and communicating with the central bore through a plurality of spill ports.
  • Each pumping means communicates with a combustion chamber of an internal combustion engine.
  • the central bore communicates with a source of fuel at a slight positive pressure.
  • the plurality of plungers are arranged to be reciprocated through the rotary motion of a swash plate actuator mechanism and the quantity of fuel to be delivered to the engine is determined by the duration of closure of the spill port through the rotary actuation of a land mechanism which is received within the central bore. Timing of fuel delivery is also determined by the shape of the land. The land is rotated in synchronism with engine operation.
  • Means are provided for altering the axial position of the rotary land within the central bore or passage and means are also provided for adjusting the rotational phasing of the land relative to rotation of the plunger actuating mechanism to achieve control of the duration and of the timing of spill port closure.
  • the shape of the rotary land and hence the fuel delivery characteristic of the pump is determined with respect to the fuel requirements of the engine.
  • FIG. 5 illustrates a typical set of air consumption curves, and hence fuel requirement curves, for an internal combustion engine having a throttled air inlet wherein the air consumption in pounds per cycle is graphed as a function of engine speed for three power levels of operation.
  • the upper curve corresponds to maximum power, or wide open throttle operation, while the two lower curves correspond to an intermediate and a low power range of operation both of which may be termed part throttle operation.
  • FIG. 6 illustrates fuel delivery curves of a prior art pump operated in conjunction with an engine having a throttled air inlet and corresponding to the air consumption curves of FIG. 5.
  • the part throttle curves illustrate that fuel delivery increases with respect to r.p.m., and the increase is more pronounced at the lighter load operation than at the intermediate loading level.
  • This increase in fuel delivery as a function of load for increasing speed at part throttle settings is a result of the fact that the swash plate mechanism causes the linear velocity of the pumping plunger to be a maximum at before top dead center (TDC) and to be gradually decreasing from that point to TDC.
  • the injection advance requires, at part throttle settings, that injection occur at high r.p.m. in proximity to 90 before TDC and at lesser degrees of advance (closer to TDC) as the speed of the engine decreases.
  • the linear velocity of the pumping plunger is greater during the delivery portion of the stroke at large degrees of advance (higher engine speeds) causing an increase in the length of the delivery portion of the plunger stroke and increased pumping efficiency at higher speeds both of which result in increased fuel delivery with increased engine speed.
  • a secondary factor which contributes to the fuel delivery characteristic increasing with speed results from the efficiency of the prior art pumps increasing with increases in engine speed. These efficiency increases are due to (1) dynamic injection effects at the beginning and ending of each injection stroke which become more significant as engine speed and hence plunger speed increases and (2) to plunger and metering sleeve leakage which decreases with increases in plunger speed.
  • the velocity of the pumping plunger increases causing pumping pressure to reach the check valve opening pressure more rapidly during the transient conditions of spill port closings and openings. Leakage around the pumping plunger is also reduced since the time period during which leakage occurs is reduced at higher speeds further improving pumping efficiency.
  • the axial and phase positioning of the spill port closing land structure is accomplished by the combination of a governor mechanism and an operator controlled positioning lever.
  • suitable linkage mechanism could be designed to sense and respond to various conditions of engine loading to modulate or otherwise override the normally provided linkage mechanism, such an approach would result in a marked increase in cost and complexity and would be subject to tolerance buildup errors and/or to a degree of insensitivity either of which would make such an approach unattractive. It is, therefore, an object of the present invention to provide a mechanism for modulating fuel delivery as a function of engine loading which does not require the addition of linkage mechanism ex ternal of the fuel pump. It is a further object of the present invention to provide a mechanism for modulating fuel delivery in response to various degrees of engine loading and wide open throttle operation which is wholly contained within the fuel injection pump.
  • a swash plate mechanism as the plunger actuator.
  • a swash plate may be visualized as a cylinder having an actuator surface defined by the intersection of a plane with the axis of the cylinder at an angle other than a right angle. The particular angle of intersection is determined by the radius of the cylinder and the desired total plunger displacement.
  • This actuator surface results in a need for fairly complex associated structure to mate with the actuator end or foot of the plunger, as for example, a ball and socket arrangement, or in the alternative, that the actuator ends of the plungers and the actuator surface must be provided with a suitable wear resistive material or must be manufactured using a surface hardening process to prevent the point contact between plunger and actuator surface from cutting a groove in the surface or otherwise abraiding the actuator surface and the actuator end or foot of the plunger. This is of course, a result of the constantly changing actuator surface angle with respect to the foot of any one of the plungers.
  • the present invention provides for a suitably contoured face cam in place of the normally provided swash plate mechanism to tailor the distance and rate of plunger displacement to more closely match the fuel delivery characteristics of the pump with fuel requirements of the engine.
  • a plunger driving face cam according to the present invention is provided with a contour which actuates the plungers to provide for a gradually increasing rate of plunger displacement over a major angular segment of the portion of the face cam preceding top dead center (TDC).
  • TDC top dead center
  • the velocity characteristics of the cam are sinusoidal at the beginning and ending of plunger lift but are flat with a gradual increase in velocity in the operating range. This permits the pump designer to provide plunger displacement as a controlled variable, along with land shape, to determine fuel delivery timing and quantity of fuel pumped.
  • the cam surface is arranged to be radial to the axis of rotation of the cam. That is, any straight line drawn from the cam surface to the axis of rotation of the cam will intersect the axis of rotation at a right angle.
  • Each plunger is provided with a cylindrical foot, with the axis of the cylinder intersecting the axis of rotation of the cam at a right angle (and therefore parallel to the actuating surface of the face cam) to provide for line contact between the cam surface and the plunger foot.
  • Means in the form of a shaped opening having a flat portion cooperating with a flat surface provided therefor on the plunger side is provided to keep the plunger foot axis in proper alignment particularly at TDC and 180 behind TDC.
  • FIG. 1 is a front elevational view of a fluid or fuel injection pump with which the present invention is of utility.
  • FIG. 2 is an enlarged cross sectional view of the pump assembly according to FIG. 1 utilizing the present invention.
  • FIG. 3 is a schematic representation of the fluid or fuel injection pump assembly shown in FIG. 2 illustrating the timing and duration control mechanism.
  • FIG. 4 is a representation of the fluid metering land employed by the fluid or fuel injection pump of FIG. 2.
  • FIG. 5 shows a graph of air consumption by an internal combustion engine having a throttled air inlet as a function of engine speed with one curve representative of air consumption at full load operation and two curves representative of air consumption at part load operation.
  • FIG. 6 shows a graph of fuel delivery as a function of engine speed for prior art injection pump for the three levels of air consumption illustrated in FIG. 5.
  • FIG. 7 shows a graph of fuel delivery as a function of engine speed for a similar injection pump which incorporates the present invention for the three levels of engine air consumption illustrated in FIG. 5.
  • FIG. 8 shows a sectional view of a plunger actuating face cam of the present invention.
  • FIG. 9 shows a sectional view taken along lines 9-9 in FIG. 2, of a pumping plunger according to the present invention.
  • FIG. 10 shows a top elevational view of the antirotation plate of one aspect of the present invention.
  • FIG. 11 is a series of graphs illustrating the operation of the prior art pumps wherein FIG. 11a illustrates the spill port opening and closing positions as functions of engine speed at a selected light load level of operation and FIG. 11b illustrates the plunger displacement and rate of displacement curves for a typical prior art swash plate actuating mechanism.
  • FIG. 12 is a series of graphs illustrating the operation of the fuel injection pump incorporating the actuating mechanism of the present invention wherein FIG. 12b illustrates the plunger displacement and rate of displacement curves for the actuating mechanism of the present invention and FIG. 12a is identical to FIG. lla.
  • FIG. 1 an external view of the fluid or fuel injection pump assembly according to the present invention which includes a housing enclosing the actuating mechanism and pump plunger for the plunger type fuel injection pump of the present invention.
  • This housing 10 has a number of fuel discharge outlets 11 that are each adapted to be connected by an external hose to an individual fuel injec' tion nozzle, not shown.
  • the housing 10 includes a window 12 for observing the degree of injection advance; that is, for observing the timing of the pump to the engine, which will be described in more detail hereinbelow. .It also includes a window 14 adjacent an internal fuel reservoir to observe for air bubbles, etc., in the fuel for fuel vapor evaluation purposes.
  • An externally controlled linkage 16 forms a portion of the engine control linkage mechanism, and partially controls the discharge of fuel through outlets 11 in a manner to be described. More specifically, as shown in FIG. 2, pump housing 10 has several sections bolted or otherwise secured together.
  • cover portion 20 includes a cover portion 20, a hollow upper housing portion 22 containing a flyweight governor mechanism, a central housing portion 26 containing the fuel pump plungers and fuel flow metering sleeve valve, and a lower housing portion 28 enclosing the drive shaft and the pump drive plate assembly according to the present invention.
  • Lower housing portion 28 has a central bore 30 in which is rotatably mounted a tubular drive shaft 32 on side bearings 33.
  • the drive for this shaft is not shown, but it would, in general be driven by a geared take-off from the engine cam 0r crank shaft.
  • the upper part of shaft 32 is formed integral with the drive plate or face cam 34 of the present invention, having a contoured drive surface 36 to be described in greater detail hereinbelow. Face cam 34 is rotatably mounted on thrust bearings 40.
  • a reciprocating-type piston or plunger 50 is axially movable in each of a plurality of bores or barrels 52 and the lower end 53 of each of these pistons or plungers 50 bears on the contoured drive surface 36 of face cam 34.
  • Anti-rotation plate 54 cooperates with flats 51 provided on pistons or plungers 50 to prevent rotation of the plungers about their own axes, particularly at the top and bottom dead center positions ofeach plunger. This will be described in greater detail hereinbelow with reference to FIGS. 9 and 10.
  • each bore or barrel 52 is intersected by cross bores 56 and 58 in housing portion 26, the number of bores or barrels 52 corresponding to the number of cylinders or combustion chambers in the engine to which the fuel pump is to be connected.
  • Bores 56 constitute discharge passages for fuel leading to delivery valves 59.
  • Passages 58 constitute spill holes or ports connected to a central bore or fluid chamber 60 in housing section 26.
  • a fuel storage area 61 is provided on one side of pump housing portion 22, and is connected to chamber 60 by intersecting passages 62 and 64.
  • An external fuel transfer pump (not shown) would be connected to passage 64.
  • each plunger 50 is fed with fuel from the chamber 60 through its associated spill holes 58 during the suction stroke or downstroke of the plunger.
  • fuel is displaced either back into chamber 60 or through the delivery valve 59.
  • Fuel under pressure on the upper end of each plunger 50 forces the lower end 53 of each of these plungers or pistons into engagement with the contoured drive surface 36 of face cam 34.
  • fuel delivery from the pump is controlled primarily by proper phasing of a sleeve valve 70, having a metering helix 72, with respect to plunger displacement.
  • the sleeve valve is capable of executing both axial and rotational movements to variably position metering helix 72 to close or open spill ports 58.
  • This sleeve valve 70 is of the spool type, and has upper and lower lands 73 and 74 connected by a necked portion 76 of reduced diameter.
  • the land diameters are such as to effectively seal the annular internal fuel reservoir defined between the lands while the diameter of the metering helix 72 is effective to seal spill ports 58 so that axial and rotary movement of the sleeve valve 70 will control fuel flow into chamber 60 from passage 62 and also the passage of fuel into and out of spill ports 58.
  • Metering helix 72 may be formed as a raised circumferentially extending portion of the lower land 74 in the shape of a helix that is integral with the valve 70 and moves axially and rotatably with it, in a manner to be described, to progressively close or open spill ports 58.
  • the metering sleeve valve 70 surrounds an extension 82 of drive shaft 32, and can be moved both axially and rotatably with respect to it in the following manner.
  • a tubular pump timing bushing is bolted to drive shaft 32 which is internally splined to the lower end of shaft 82.
  • the two shafts are further connected by a pin 84 that is fixed to bushing 80 and projects through a slot 86 in shaft 82 to permit relative axial movement between them.
  • Shaft 82 projects through an oil seal 87 and a thrust bearing 88, the thrust bearing being located axially in one direction by a retaining ring 90.
  • Shaft 82 continues upwardly freely through metering valve 70, through a governor assembly 91, and into cover portion 20 through a journal bearing 92.
  • the bearing 92 is inserted in an aperture in a partition 93 in housing portion 22.
  • the upper end of shaft 82 has a slotted pivotal connection at 94 to a lever 95 that is fixed to a rotatably mounted rod 96 forming a part of the vehicle control linkage.
  • Thrust bearing 88 and metering valve 70 are axially separated by control spring 97 which surrounds shaft 82.
  • the spring is seated at one end in a recess 98 in valve land 74 and at the other end against thrust bearing 88.
  • the spring exerts a predetermined upward preload on metering valve 70.
  • the upper end of valve 70 is also recessed, and is slidably (see FIG. 3) splined to a reduced diameter sleeve member 100.
  • the sleeve member is fixed to the lower race 102 of a thrust bearing 104 that slidably surrounds shaft 82.
  • Sleeve member 100 has a cam follower slot 106 in which slides a drive pin 108 that is fixed to shaft 82.
  • Slot 106 has an axially extending portion 110 and an inclined portion 1 12.
  • Axial portion 110 permits relative axial movement between metering valve 70 and shaft 82, while inclined portion 112 forces the metering valve 70 to rotate to shaft 82 when the valve is moved axially.
  • a second compression spring 114 surrounds shaft 82, and is seated between the lower end of sleeve member 100 and the bottom of recess 115 in upper land 73.
  • the thrust bearing 104, sleeve member 100, and valve 70 are moved axially downwardly by the mechanical flyweight governor assembly 91.
  • This latter mechanism includes a cage or base plate 118 nonrotatably keyed to shaft 82.
  • the cage has pairs of laterally spaced arms or ears 120, between which are pivotally mounted a pair of right-angled speed responsive members 122.
  • the lower portion of each member 122 is formed as a weight 124, while the upper portion constitutes a lever 125 that abuts the upper race of bearing 104.
  • Suitable screw adjusting devices 126 (in FIG. 2) provide adjustment in a known manner.
  • the governor operates to depress sleeve valve 70 downwardly relative to shaft 82 against the forces of springs 97 and 114 upon outward movement of weights 124 under the effect of centrifugal force, in a manner further explained hereinbelow.
  • the preload of spring 114 maintains sleeve member 100 and metering valve 70 in their axially most separated positions shown in FIGS. 2 and 3 below a predetermined speed of rotation of shaft 82 of, say for example, 2,400 rpm; that is, below the speed at which injection advance is desired, as will be explained more fully hereinbelow. Above the speed at which injection advance is desired, centrifugal force acting on the governor weights 124 overcomes the preload of spring 114, and permits movement of sleeve 100 into the end of valve land 73.
  • Lubrication of the various parts is as follows:
  • the accelerator control linkage and governor mechanism, as well as the upper portion of metering valve 70, are Iubricated by fuel sprayed into the upper housing portion chamber by means of a nonreturn lubricating jet assembly indicated at 127.
  • the pump drive plate thrust and side bearings 40 and 33, and plungers 50, are lubricated by engine oil supplied through suitable intersecting passages 127' leading to these parts.
  • the accelerator pedal control linkage is illustrated schematically in FIG. 3. It includes an accelerator pedal 128 pivotally mounted at 129 and pivotally connected near its center to an articulated linkage consisting of links 130 and 131. These links 130 and 131 are pivotally connected to each other, with the opposite end of link 131 being fixed to the throttle valve 132 to control the quantity of air flowing into the intake manifold 133.
  • Conduit 134 communicates the pressure within intake manifold 133 to a vacuum motor 135.
  • Movable wall or diaphragm 136 is connected to linkage member 137 which is pivotally connected to linkage member 138.
  • the other end of linkage member 138 is fixed to rod 96. Suitable wide open throttle and idle stops 139a and 139 are provided, as shown.
  • a return spring 140 normally biases the pedal 128 to its idle position.
  • an additional external fuel shut-off linkage is provided. This consists of a power or manually movable knob 141 securedby a horizontally movable link 142.
  • the lever in its simplest form, is fixed for rotation with rod 96 as shown in FIG. 3, and rotates between fuel shut-off and wide-open throttle positions, as indicated.
  • the fuel shut-off position would move the metering helix 72 downwardly to a position completely opening spill ports 58 so that no fuel would be discharged through passages 56.
  • FIG. 2 shows the parts of the fluid or fuel injection pump assembly of the present invention described thus far in the curb stop or nonrunning position.
  • the idle speed and injection advance springs 97 and 114 preload metering valve 70, sleeve member 100, and the governor members 122 to the positions shown.
  • the manifold vacuum is maximized and diaphragm 136 is at its rightward (relative to FIG. 3) extreme position; therefore, no upward force is exerted on shaft 82 by this linkage.
  • the metering helix 72 is positioned relative to the spill holes 58 at the point indicated in FIG. 4 as the curb stop position so that slightly more than the normal amount of fuel required to provide idle operation of the engine would be injected past the delivery valves if the fuel pump were to be driven at this time.
  • auxiliary fuel control shut-off knob 141 (FIg. 3) would be moved to the left to move shaft 82 and valve downwardly to a fuel shut off position locating the spill holes 58 relative to helix 72 so that all of the fuel will spill back into chamber 60. This would correspond to the upper point 78 of the helix 72 being axially below the spill ports 58.
  • the accelerator pedal 128 is depressed fully to its wide open throttle position, which also corresponds to the prime-start position. This full depression of the pedal causes lever to rotate clockwise to move shaft 82, governor mechanism 91, and metering valve 70 upwardly so that the bottom portion of the metering helix 72 now closes off spill holes 58 for almost the entire rotation of the metering valve as indicated in FIG. 4 as the prime start position.
  • drive pin 108 will be positioned in the slot 106 of sleeve 100 at the junction between portions 110 and 112. Further change in the axial position of metering helix 78 between 900 and 2,400 rpm. is now,
  • Deceleration control is obtained by releasing the accelerator pedal to its idle position. Shaft 82, the governor assembly 91, and helix 72 immediately move downwardly, and decrease the fuel output. Since the governor is operative, the high centrifugal force still acting on the governor weights at first maintains spring 114 compressed, and the sleeve 100 almost entirely within land 72. However, as the speed decreases, the force of spring 114 will move valve 70 downwardly, so that at 2,400 rpm, spring 114 will have moved valve 70 to its downwardmost position relative to shaft 82, and valve 70 will be in its downwardmost position against stop 99. The helix 72 will now completely uncover the area of spill holes 58, and thereby shut off all fuel flow to the nozzles. When the speed falls within the 400-900 r.p.m. idle speed range, spring 97 and the governor will again be operative to move the sleeve valve 70 and metering helix 72 to the idle speed position.
  • FIGS. 5 and 6 a series of graphs illustrating air and fuel consumption for a typical internal combustion engine are shown.
  • FIG. 5 illustrates the air consumption in pounds per cycle graphed as a func- 204 illustrates the air consumption at a maximum power operation or wide open throttle setting and it can be seen by inspection of this curve that at wide open throttle, the air consumption does change as a function of engine speed with a significant decrease in the rate of consumption in the high rpm. range.
  • FIG. 6 illustrates fuel delivery at three operational loadings or throttle settings which correspond to those illustrated in the FIG. 5 graph.
  • the curves 206 and 208 demonstrate a fuel delivery characteristic which increases with increasing r.p.m.
  • the curve 210 illustrates a fuel delivery characteristic which very closely matches the air consumption characteristic curve 204 from the graph of FIG. 5.
  • the configuration of the metering land or helix and its axial and rotary phasing were arranged to provide for a close fit between the fuel delivery characteristic and the air consumption characteristics at wide open throttle settings. This necessarily results in the pumping characteristic being excessive at operational loadings less than maximum or at part throttle operation. 1
  • FIG. 11a a series of graphs are shown illustrating by the solid lines, the angular position of the metering helix 72, and consequently the opening and closing events and duration of closure of the spill port 58, relative to plunger top dead center for a representative pumping means as a function of engine speed for a selected light load, or part throttle, operation for a fuel injection pump associated with a representative internal combustion engine having a throttled air inlet.
  • Angular position is expressed in degrees before pumping plunger TDC. It can be seen from the FIG.
  • FIG. 11a also illustrates by dashed lines, the spill port closing and opening events and duration of spill port closure at a high load or wide open throttle operation over the expected range of engine speeds.
  • the total potential spill port closure is approximately 107 of angular rotation extending from about 146 before plunger TDC to about 40 before plunger TDC for this particular engine and prior art pump -combination.
  • the injection duration and injection advance can be directly controlled by the axial and rotary positioning of the metering land or helix 72 relative to spill port 58 by the mechanism described with reference to FIGS. 2 and 3 to suit the engine fuel delivery requirements.
  • the plunger displacement and rate of displacement curves are graphed as a function of the angular positioning of a typical prior art swash plate type actuating mechanism.
  • the angular positioning of the swash plate has been illustrated in alignment with the angular positioning for the spill port closing and opening events.
  • the dynamic pumping effects associated with the differences in plunger velocity are minimal or zero, that is, that the opening and closing events are instantaneous and leakage is constant or zero
  • the pumping mechanism as a positive displacement pump which will pump fuel to the engine only during the period of spill port closure.
  • FIG. 11b graph illustrates a plunger displacement of approximately 5.4 units of distance for the low speed spill port closure (which occurs without any additional amount of injection advance) as compared with a plunger displacement of approximately 6.7 units of distance for the high speed spill port closure event (which occurs at a maximum condition of injection advance).
  • the rate of displacement curve is one-half of a sinusoidal wave which reaches a maximum value at a point 90 before plunger TDC and thereafter decreases to a value of zero at plunger TDC.
  • the plunger is moving with a relatively higher velocity at the position of maximum injection advance (at the illustrated part throttle operation) and is moving with a much lower relative velocity at the minimum degree of injection advance.
  • the plunger velocity and therefore the distance of plunger displacement is greatly different for different amount of injection advance. This difference can also be seen to contribute to increases in fuel pumped for increasing engine speed under constant load operation of the associated engine over the speed range of operation of that engine.
  • the plunger displacement curve of FIG. 11b will be recognized as constituting the profile curve of the actuating mechanism applied to the plunger, expressed in terms of axial height from a zero reference for degrees of rotation before plunger TDC with the maximum displacement corresponding to the plunger TDC and identified as zero degrees.
  • This curve is recognized as being one half of a sinusoidal wave going from a minimum value at 180 of rotation before plunger TDC and going to a maximum value at plunger TDC.
  • This curve is the displacement curve generated by a typical swash plate actuating mechanism as defined hereinabove.
  • FIGS. 8 and 12 the actuating mechanism of the present invention will be described and is illustrated. With particular reference to FIG. 8,
  • the actuating mechanism 34 of the present invention is provided with an actuating surface 36 from which an imaginary line drawn to intersect the axis of rotation will intersect that axis at a predetermined, constant angle without regard to the location on the actuating surface from which the line is drawn.
  • this angle of intersection is a right angle so that the actuating surface may be termed to be radially directed with respect to the axis of rotation of the actuating mechanism 34.
  • This actuating surface 36 constitutes a ramp surface for actuating the pumping plungers of the fuel injection pump.
  • the circumferential profile of actuating surface 36 is arranged to provide a pumping contour in terms of linear displacement for varying degrees of angular rotation which differs significantly from that illustrated by the displacement curve of FIG. 11b.
  • FIG. 12 wherein FIG. 12a is substantially identical with FIG. 11a and with particular reference to FIG. 12b it can be seen that the differences between the displacement curves of the prior art device and of the instant device superficially appear to be similar. However, a consideration of the velocity or rate of displacement curves quickly disspells any such thoughts of similarity.
  • the displacement curve of FIG. 126 has segments a, b and c all of which appear to be linear. Segments a and c are, in fact, short segments of larger sinusoidal curves whose importance is to provide a rapid initial acceleration and a rapid terminal deceleration for the plungers. Segment b of this curve is of cardinal importance. Segment b of the rate of displacement curve of FIG. 12b is provided with a slope which does not decrease in the region between before TDC and TDC but which increases over the range of angular rotation from about l47 /2 before plunger TDC, through 90 before TDC to about 42 /z before plunger TDC.
  • the undesired efficiency improvement effects at high engine speeds as described hereinabove can be largely avoided or eliminated by providing a slight decrease in plunger velocity and hence slightly lower displacement at higher engine speeds.
  • this is not apparent due to the fact that the plunger velocity curve reaches a maximum value slightly in advance of the opening of the spill port and termination of the spill port closure event and that thereafter the plunger is rapidly decelerated according to the profile of segment c.
  • the plunger has been decelerated and hence executes somewhat less displacement motion and therefore pumps a slightly lesser quantity of fuel.
  • plunger displacement curve of FIG. 12b corresponds to the circumferential profile curve of the actuating surface 36 expressed in terms of linear displacement in the axial direction from a reference for degrees of angular rotation of that actuating surface with respect to TDC.
  • the actuating mechanism 34 may be radially ground on suitable radial grinding machinery which may be manually or automatically controlled to provide for the angular profile curve expressed as the displacement curve of FIG. 1212.
  • the actuating surface 36 is arranged to be radially directed, or perpendicular to the axis of rotation of the actuating mechanism 34. This permits the pump manufacturer to ignore any displacement or velocity changes in the plunger motion occasioned by changes in the angular relationship between the feet of the plungers and the actuating surface 36.
  • the foot 53 of the plunger 50 according to FIG. 9 is arranged to be cylindrical with the axis of the cylinder being directed parallel to the actuating surface 36.
  • the axis of the cylinder is directed perpendicularly to the axis of rotation of the actuating mechanism 34 so that the foot of the plunger will be in line contact with the actuating surface 36 for all positions of angular rotation. Since the line contact will coincide with the axis of the plunger 50 at the TDC and at 180 before TDC positions, undesirable rotation of the plunger about its own axis may readily occur. A partial rotation of this plunger about its axis could readily result in a point contact occurring between an edge of the foot 53 of the plunger 50 and actuating surface 36. This point of contact would rapidly wear away or otherwise abraid both the plunger 50 and the actuating surface 36. To
  • holes 55 equal in number to the number of plungers 50 and each such hole 55 in this instance is provided with straight edged region 55'. These straight regions 55' are sized and positioned to cooperate with flats provided therefor on the sides of pumping plungers 50 to prevent any rotation of these plungers about their own axis.
  • actuating surface 36 for the actuating mechanism 34 which has a contour shaped to provide a desired displacement and velocity of displacement curve the pumping efficiency effect and the velocity change effect occasioned by prior art devices which made their fuel pumping characteristic different from the air consumption characteristic of their associated engine can be readily avoided.
  • the particular contour applied to actuating surface 36 may be readily selected to conveniently match the air consumption characteristic of the associated engine.
  • the present invention simplifies the required structural elements of a fuel injection pump which may be used and also assures that the velocity and rate of plunger displacement will not be deleteriously affected by changing angular relationships between the actuating surface and the feet of the plunger 53.
  • the fuel injection pump according to the present invention provides fuel delivery at part throttle operation, illustrated by curves 212, 214 which closely matches air consumption as illustrated by curves 200, 202 of FIG. 5.
  • the full load or wide open throttle fuel delivery curve 216 also closely matches the air consumption curve 204 of FIG. 5.
  • a fuel injection pump for an air throttled internal combustion engine comprising in combination:
  • a metering valve means arranged to close the spill passage means in a selected sequence
  • drive means adapted to be driven by the engine and including means for driving said metering valve to thereby sequentially close said spill passage means in the selected sequence and means for reciprocating said plungers within said bores;
  • said engine responsive means including speed responsive means within said housing operatively coupled to said metering valve and arranged to controllably vary the phase relationship between spill passage means closure events and engine operation whereby the spill passage means closure event may be advanced with increasing engine speed over selected ranges of engine operation;
  • said means for reciprocating the plungers including rotary drive means arranged to engage said plungers, operative upon rotation to reciprocate said plungers within said bores;
  • said rotary drive means including a contoured cam actuating surface for engagement with said plungers, the contour of said cam actuating surface being selected to provide plunger displacement velocities which do not decrease for drive means rotational positions corresponding to substantially all possible rotational positions of spill port closure by said metering valve means, whereby the cam actuating surface cooperates with the advance of the metering valve means and the concomitant advance of the spill passage means closure events to provide a fuel delivery characteristic which closely matches the air flow characteristic of the engine over substantially the entire speed range of the engine.
  • cam actuating surface is contoured to provide for an increasing velocity of plunger displacement for a major portion of the angular rotation of said rotary drive means prior to attainment of plunger top dead center within the plunger bore.
  • cam actuating surface is arranged to provide for a maximum velocity of plunger displacement at an angular position substantially corresponding to the angular position of the spill port closing event most closely approaching plunger top dead center.
  • cam actuating surface is arranged to provide for a velocity of plunger displacement which is increasing for substantially all of the cam means angular rotation corresponding to the total potential spill port closure.
  • a stationary housing having a central bore and a source of fluid under pressure connected to said bore;
  • drive shaft driven fluid pumping means having a plurality of fluid discharge passages each connected in parallel to a fluid discharge injection line and to a fluid spill line communicating with said bore, said spill line receiving fluid from and spilling fluid into said bore;
  • a drive shaft driven fluid metering valve member slidably and rotatably mounted in said bore between said source and said spill lines;
  • said metering member having fluid flow control portions thereon variably controlling the flow of fluid through said spill lines as a function of the axial and rotative movements of said metering member;
  • said limited movement connection between said metering and drive shaft members including a cam element and cam follower element operably engaged and connected respectively one element to each of the said metering and shaft members whereby rotation of said drive shaft member above a first predetermined speed effects a relative rotation between said members to vary the scheduling of control of said spill lines by said metering member;
  • said drive shaft driven fluid pumping means including a plurality of piston members arranged for reciprocation within a plurality of pumping cylinders and a rotary piston actuator having contoured drive surface, said drive surface having a fluid pumping segment for providing a piston displacement velocity which does not decrease for rotation of the drive surface through angular positions of rotation corresponding with substantially all possible scheduling of the control of the spill lines by said metering member whereby said fluid pumping segment and said limited motion connection may be made cooperative to provide a rate of fluid pumping substantially proportional to the rate of air consumption by the engine.
  • a fuel injection pump assembly for an air-throttled internal combustion engine comprising, in combination:
  • drive shaft driven fluid pumping means arranged to be driven by the engine in relation to engine operational events and having a plurality of fluid spill lines connected at one end to discharge outlets and at their other ends to said bore to receive fluid from said source and to spill fluid into said bore;
  • valve member slidably and rotatably mounted in said bore surrounding said drive shaft member and having a lost motion connection thereto;
  • valve member having a fluid metering helix portion thereon variably closing said spill lines as a function of the axial and rotative movements of said valve member to control flow volume of fluid to said outlets;
  • movable external control means connected to said drive member for moving it and said valve member axially;
  • said drive shaft driven fluid pumping means further including a plurality of pumping cylinders having plungers disposed therein and a rotary actuator driven in synchronism with said drive member;
  • said rotary actuator having an actuating surface arranged to reciprocate said plungers within their cylinders and contoured to provide a velocity of plunger displacement which does not decrease for substantially all possible spill port closure events whereby said fluid pumping segment and said limited motion connection may be made cooperative to provide a rate of fluid pumping substantially proportional to the rate of air consumption by the engine.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Fuel-Injection Apparatus (AREA)
  • High-Pressure Fuel Injection Pump Control (AREA)
US00318297A 1972-12-26 1972-12-26 Fuel injection pump Expired - Lifetime US3856438A (en)

Priority Applications (5)

Application Number Priority Date Filing Date Title
US00318297A US3856438A (en) 1972-12-26 1972-12-26 Fuel injection pump
CA181,381A CA987560A (en) 1972-12-26 1973-09-18 Fuel injection pump for an internal combustion engine
GB5610673A GB1425875A (en) 1972-12-26 1973-12-04 Fuel injection pump for an internal combustion engine
JP13746573A JPS5425569B2 (pl) 1972-12-26 1973-12-11
DE2362526A DE2362526A1 (de) 1972-12-26 1973-12-15 Brennstoff-einspritzpumpe fuer brennkraftmaschinen

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
US00318297A US3856438A (en) 1972-12-26 1972-12-26 Fuel injection pump

Publications (1)

Publication Number Publication Date
US3856438A true US3856438A (en) 1974-12-24

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US00318297A Expired - Lifetime US3856438A (en) 1972-12-26 1972-12-26 Fuel injection pump

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US (1) US3856438A (pl)
JP (1) JPS5425569B2 (pl)
CA (1) CA987560A (pl)
DE (1) DE2362526A1 (pl)
GB (1) GB1425875A (pl)

Cited By (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3986487A (en) * 1973-08-29 1976-10-19 Nissan Motor Co., Ltd. Rotary type fuel injection pump
EP0007798A1 (en) * 1978-07-26 1980-02-06 Ford Motor Company Limited Fuel injection pump
US4301777A (en) * 1979-11-28 1981-11-24 General Motors Corporation Fuel injection pump
US4530331A (en) * 1984-03-27 1985-07-23 Caterpillar Tractor Co. Thrust and planetary gear coupling for a rotor of a distributor fuel injection pump
US5931644A (en) * 1995-03-30 1999-08-03 Caterpillar Inc. Precision demand axial piston pump with spring bias means for reducing cavitation
US20150316045A1 (en) * 2012-12-10 2015-11-05 Kongsberg Automotive Ab Unitary Fluid Flow Apparatus for Inflating and Deflating a Device
US11401883B2 (en) * 2020-04-03 2022-08-02 Ford Global Technologies, Llc System and method for direct injection fuel pump control

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE2907279A1 (de) * 1979-02-24 1980-08-28 Inst Motorenbau Prof Huber E V Kraftstoffeinspritzsystem fuer verbrennungsmotoren
DE2909307A1 (de) * 1979-03-09 1980-09-18 Inst Motorenbau Prof Huber E V Kraftstoff-einspritzsystem mit kontinuierlicher foerderung und intermittierender einspritzung

Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2061144A (en) * 1932-04-12 1936-11-17 Stoutz Robert De Piston spinning pump
US2775210A (en) * 1951-04-10 1956-12-25 Scintilla Ltd Piston pump
US2882831A (en) * 1954-06-17 1959-04-21 Gen Electric Constant flow positive displacement mechanical hydraulic unit
US2992619A (en) * 1950-08-05 1961-07-18 William C Nilges Fluid pumps, motors and methods therefor
US3045604A (en) * 1960-05-04 1962-07-24 Fmc Corp Multi-cylinder pump
US3046950A (en) * 1958-01-22 1962-07-31 Whiting Corp Constant mechanical advantage rotary hydraulic device
US3319568A (en) * 1965-07-16 1967-05-16 Ford Motor Co Fuel injection pump assembly
DE1914598A1 (de) * 1968-03-22 1970-08-27 Nat Res Dev Hydrostatische Kolbenmaschine

Patent Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2061144A (en) * 1932-04-12 1936-11-17 Stoutz Robert De Piston spinning pump
US2992619A (en) * 1950-08-05 1961-07-18 William C Nilges Fluid pumps, motors and methods therefor
US2775210A (en) * 1951-04-10 1956-12-25 Scintilla Ltd Piston pump
US2882831A (en) * 1954-06-17 1959-04-21 Gen Electric Constant flow positive displacement mechanical hydraulic unit
US3046950A (en) * 1958-01-22 1962-07-31 Whiting Corp Constant mechanical advantage rotary hydraulic device
US3045604A (en) * 1960-05-04 1962-07-24 Fmc Corp Multi-cylinder pump
US3319568A (en) * 1965-07-16 1967-05-16 Ford Motor Co Fuel injection pump assembly
DE1914598A1 (de) * 1968-03-22 1970-08-27 Nat Res Dev Hydrostatische Kolbenmaschine

Cited By (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3986487A (en) * 1973-08-29 1976-10-19 Nissan Motor Co., Ltd. Rotary type fuel injection pump
EP0007798A1 (en) * 1978-07-26 1980-02-06 Ford Motor Company Limited Fuel injection pump
US4301777A (en) * 1979-11-28 1981-11-24 General Motors Corporation Fuel injection pump
US4530331A (en) * 1984-03-27 1985-07-23 Caterpillar Tractor Co. Thrust and planetary gear coupling for a rotor of a distributor fuel injection pump
US5931644A (en) * 1995-03-30 1999-08-03 Caterpillar Inc. Precision demand axial piston pump with spring bias means for reducing cavitation
US20150316045A1 (en) * 2012-12-10 2015-11-05 Kongsberg Automotive Ab Unitary Fluid Flow Apparatus for Inflating and Deflating a Device
US10107279B2 (en) * 2012-12-10 2018-10-23 Kongsberg Automotive Ab Unitary fluid flow apparatus for inflating and deflating a device
US11401883B2 (en) * 2020-04-03 2022-08-02 Ford Global Technologies, Llc System and method for direct injection fuel pump control

Also Published As

Publication number Publication date
JPS4989017A (pl) 1974-08-26
DE2362526A1 (de) 1974-06-27
JPS5425569B2 (pl) 1979-08-29
GB1425875A (en) 1976-02-18
CA987560A (en) 1976-04-20

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