EP2796705A1

EP2796705A1 - Fuel injection system and fuel pump

Info

Publication number: EP2796705A1
Application number: EP13164627.5A
Authority: EP
Inventors: Philip Dingle
Original assignee: Delphi International Operations Luxembourg SARL
Current assignee: Delphi International Operations Luxembourg SARL
Priority date: 2013-04-22
Filing date: 2013-04-22
Publication date: 2014-10-29

Abstract

A fuel injection system (100) for an internal combustion engine is disclosed. The fuel injection system comprises at least one electronically-controlled fuel injector (102), a fuel pump (104) for supplying pressurised fuel to the or each injector (102), and a controller (130) arranged to control the injection of fuel from the or each injector (102). The fuel pump (104) comprises a pump chamber (146), a pumping element (140), and a drive mechanism (360) for driving the pumping element (140) in a pumping stroke in which the volume of the pump chamber (146) is reduced, and a filling stroke in which the volume of the pump chamber (146) is increased and in which fuel is admitted to the pump chamber (146) from a fuel source (116). The drive mechanism (360) comprises a biasing arrangement (468) arranged to apply a resilient biasing force, such as a spring force, to the pumping element (140) to drive the pumping stroke. In this way, the resilient biasing force acts to pressurise the fuel for injection, providing a simplified fuel injection system with electronic control of the injection timing and duration. Fuel pumps suitable for use in such a system are also disclosed.

Description

Field of the invention

The present invention relates to a fuel injection system and a fuel pump suitable for use in such a system. In particular, but not exclusively, the invention relates to a fuel injection system for a compression-ignition internal combustion engine.

Background to the invention

Figure 1 of the accompanying drawings shows a known electronically-controlled fuel injection system 10 for a compression-ignition internal combustion engine. The fuel injection system 10 comprises a plurality of fuel injectors 12 (in this example, two fuel injectors 12 are shown), which are supplied with fuel at a relatively high pressure from an accumulator volume in the form of a fuel rail 14. The fuel rail 14 is connected to a high-pressure fuel pump 16, which receives fuel from a fuel tank 18 at a relatively low pressure. The high-pressure fuel pump 16 elevates the pressure of the fuel to maintain the relatively high fuel pressure in the fuel rail 14.
The high-pressure fuel pump 16 is a so-called "unit pump" of the general type described in the Applicant's granted European Patent No. EP 1 651 863 B . The pump 16 includes a pumping plunger 20 which is driven in reciprocal linear movement by a drive mechanism 22. The drive mechanism 22 comprises an engine-driven cam 24, a roller-type cam follower arrangement 26, and a return spring 28 that keeps the cam follower arrangement 26 in contact with the cam 24. The pumping plunger is 20 is connected to the cam follower arrangement 26 so that, as the cam 24 rotates, the plunger 20 is forced further into the body of the pump 16 in a forward or pumping stroke, thereby to pressurise fuel in a pumping chamber (not shown) of the pump 16, and to force the pressurised fuel from an outlet of the pump 16 to the fuel rail 14. Upon further rotation of the cam 24, the return spring 28 serves to retract the plunger 20, in a return or filling stroke, so that the pumping chamber can be re-filled with low-pressure fuel from the tank 18.
The pressure of fuel in the fuel rail 14, and hence the pressure at which fuel is injected from the fuel injectors 12, is controlled by an electronic control unit 30. The electronic control unit 30 receives signals from a plurality of sensors that allow the electronic control unit 30 to determine the appropriate target rail pressure for a given engine operating condition. In this example, the sensors comprise a throttle position sensor 32, an engine speed sensor 34, an engine phase sensor 36, and a temperature sensor 38. The electronic control unit 30 monitors the fuel pressure in the fuel rail 14 by means of a rail pressure sensor 40.
The volume of high-pressure fuel that reaches the fuel rail 14 is controlled by an inlet metering valve 42 which regulates the flow of fuel from the tank 18 to the high-pressure fuel pump 16. In some arrangements, such as that described in EP 1 651 863 B , the inlet metering valve 42 is integrally housed with the high-pressure fuel pump 16, although it is shown as a separate component in Figure 1 for clarity.
The inlet metering valve 42 is configured to close the flow path between the tank 18 and the high-pressure fuel pump 16 over a variable portion of the return stroke of the plunger 20, thereby to control the volume of fuel that is pressurised by the pump 16 in the subsequent forward stroke of the plunger 20. When the fuel pressure in the fuel rail 14 is lower than the target rail pressure, the electronic control unit 30 opens the inlet metering valve 42 over a relatively large portion of the return stroke of the plunger 20, and when the fuel pressure in the fuel rail 14 is higher than the target rail pressure, the electronic control unit 30 reduces the opening time of the inlet metering valve 42 to reduce fuel flow to the high-pressure fuel pump 16.
The flow path for fuel between the tank 18 and the high-pressure fuel pump 16 includes a transfer pump 44 and a fuel filter 46. The transfer pump 44 is typically electrically operated and runs at a constant speed whilst the engine is operating. A pressure relief valve 48 is connected to the fuel line between the transfer pump 44 and the inlet metering valve 42 to feed excess fuel back to the tank 18. The outlet of the pressure relief valve 48 is connected to a venturi 50, which provides a low-pressure drain 52 to which the injectors 12 are connected.
Injection of fuel from each fuel injector 12 is also under the control of the electronic control unit 30. As is known in the art, each injector 12 includes a control valve (not shown) which controls the movement of a valve needle (not shown) with respect to a valve seat at the tip of the injector 12. The control valve is operable to increase or decrease the fuel pressure in a control chamber (not shown) disposed at the end of the valve needle opposite the tip, by controlling flow between the control chamber and the low-pressure drain 52.
By virtue of these features, the known fuel injection system 10 can be operated to deliver fuel from each injector 12 in well-controlled and accurate injection events. The injection pressure and the injection duration can be closely controlled so that a precise volume of fuel is delivered with each injection, and so that the injection volume can be targeted to provide optimum performance and minimum emissions output based on the engine operating conditions and the throttle demand.
Fuel injection systems of the type shown in Figure 1 are therefore commonplace in sophisticated diesel engines used in, for example, automotive and high-power applications, in which performance and emissions control are sufficiently important that the relatively high cost and complexity of such a fuel injection system can be justified.
In lower-power, "nonroad" engines, for example in engines for light industrial, agricultural and construction use, with power outputs below about 30 kW (40 horse power), the demands for performance and emissions control have not traditionally been sufficient to justify the expense and complexity of an electronic fuel injection system of the type shown in Figure 1. Instead, simple mechanical fuel injection systems have generally been found to be adequate for such engines.
A mechanical fuel injection system typically comprises an engine-driven fuel pump which delivers fuel, by way of a fuel line, to an associated mechanical fuel injector. Such systems may therefore be referred to as pump-line-nozzle systems. The mechanical fuel injector includes a simple needle valve arrangement configured to open when the pressure of fuel supplied to the injector increases above a threshold value. The fuel pump, sometimes referred to as a 'jerk' pump, is arranged so that its pumping strokes are timed to deliver a burst of high-pressure fuel sufficient to cause the connected fuel injector to discharge a predetermined quantity of fuel in the appropriate part of the combustion cycle of the associated cylinder. The quantity of fuel injected with each injection is determined by the effective displacement of the pump, which is adjusted by a mechanical mechanism according to the engine operating conditions.
Mechanical fuel injection systems therefore provide a low-cost fuel injection system which offers sufficient performance for simple engines without the need for electronic control. It will be appreciated, however, that mechanical fuel injection systems offer relatively crude control of the injection pressure and injection volume. Accordingly, it is difficult to optimise the performance and emissions behaviour of engines using such systems.
Light industrial engines are, however, becoming subject to increasingly stringent emissions control legislation, which can be difficult to satisfy using mechanical fuel injection systems.
For example, in a mechanical pump-line-nozzle fuel injection system, the injection pressure follows a triangular profile, with a peak pressure of typically around 500 bar. The mean effective injection pressure for such a system is therefore of the order of 250 bar. In comparison, the mean effective injection pressure in an electronically-controlled injection system, such as shown in Figure 1, is typically 1000 to 2000 bar or even higher, and the injection pressure is relatively constant over the duration of the injection. This increased injection pressure contributes significantly to achieving improved combustion efficiency and lower emissions, but it is not accessible in known low-cost mechanical systems.
Against this background, it would be desirable to provide a fuel injection system that offers substantially improved injection control and substantially reduced emissions compared to known mechanical fuel injection systems, but which is considerably less costly and complex than known electronically-controlled fuel injection systems.

Summary of the invention

In a first aspect, the present invention resides in a fuel injection system for an internal combustion engine, comprising at least one electronically-controlled fuel injector, a fuel pump for supplying pressurised fuel to the or each injector, and a controller arranged to control the injection of fuel from the or each injector. The fuel pump comprises a pump chamber, a pumping element, and a drive mechanism for moving the pumping element in a pumping stroke in which the volume of the pump chamber is reduced, and a filling stroke in which the volume of the pump chamber is increased and in which fuel is admitted to the pump chamber from a fuel source.
The drive mechanism comprises a biasing arrangement arranged to apply a resilient biasing force to the pumping element to drive the pumping stroke. By driving the pumping stroke of the pumping element using a resilient biasing force, such as a spring force, rather than a cam arrangement, the pressure of fuel supplied to the injectors is reasonably constant and well-defined, and it is not necessary to provide a pressure sensor, control valves or other components for pressure control in the fuel injection system. However, accurate pressure-time metering of the injected fuel volume is still possible. The present invention can therefore achieve significantly reduced emissions compared to fully-mechanical fuel injection systems, but with substantially lower cost and complexity than known electronically-controlled fuel injection systems.
The pressure of fuel supplied to the or each injector is preferably determined by the resilient biasing force applied to the pumping element by the biasing arrangement. In this way, the pressure of fuel supplied to the or each injector is a pre-determined parameter of the system, rather than an operating variable. Accordingly, the controller can be configured to store a pre-determined fuel supply pressure based on the resilient biasing force, and to calculate an injection duration based on the pre-determined fuel supply pressure.
The resilient biasing force applied to the pumping element by the biasing arrangement is preferably non-linear as a function of deflection of the biasing arrangement at least over the range of movement of the pumping element. In a preferred embodiment, the resilient biasing force applied to the pumping element by the biasing arrangement is substantially constant over the range of movement of the pumping element. In this way, the variation in the pressure of fuel supplied to the or each injector can be minimised, and the fuel pressure remains relatively stable throughout the duration of each injection event.
In one embodiment, the biasing arrangement comprises a spring, and the resilient biasing force is a spring force. The spring may be a compression spring, a diaphragm spring, a Belleville washer stack spring, or another suitable type of spring. The spring may comprise a substantially constant-force spring.
When the fuel injection system includes two or more fuel injectors, each injector may be connected to a common rail that receives high-pressure fuel from the fuel pump. The common rail provides an accumulator volume for the high-pressure fuel. Alternatively, the or each injector could be connected directly to the fuel pump, without a common rail.
An accumulator volume for high-pressure fuel could be provided within each injector, in which case the injectors could be connected together in series.
In a second aspect, the present invention extends to a fuel pump for a fuel injection system. The fuel pump comprises a pump chamber having an inlet for receiving fuel at relatively low pressure from a fuel source, and an outlet for delivering pressurised fuel to the fuel injection system, a pumping element arranged for reciprocal movement with respect to the pump chamber, and a drive mechanism arranged to move the pumping element in a pumping cycle comprising a pumping stroke in which the volume of the pump chamber is reduced, and a filling stroke in which the volume of the pump chamber is increased and in which fuel is admitted to the pump chamber from the fuel source.
The drive mechanism comprises a biasing arrangement arranged to apply a resilient biasing force to the pumping element to drive the pumping stroke, and an engine-driven return mechanism arranged to counteract the resilient biasing force, thereby to enable the pumping element to undergo the filling stroke.
In contrast to typical pumps for fuel injection systems, in this second aspect of the invention the pumping stroke of the pumping element is driven by a resilient biasing force, and the return stroke is motivated by an engine-driven mechanism. This arrangement means that, advantageously, the output pressure of the pump is determined by the resilient biasing force alone, removing the need for a sophisticated pressure control system.
The resilient biasing force applied to the pumping element by the biasing arrangement is preferably non-linear as a function of deflection of the biasing arrangement over the range of movement of the pumping element. In a preferred embodiment, the resilient biasing force applied to the pumping element by the biasing arrangement is substantially constant over the range of movement of the pumping element.
The biasing arrangement may comprise a resilient biasing element, such as a spring. Thus the resilient biasing force may be a spring force. In one embodiment, the biasing arrangement comprises a compression spring. For example, the spring may comprise a conical compression spring with a variable pitch or wire diameter, so that the force-deflection behaviour of the spring is substantially flat over the range of movement of the pumping element. In another example, the biasing arrangement comprises a diaphragm spring. In a further embodiment, the pressurising spring comprises a stack of Belleville washers, preferably with an appropriate configuration to provide flat force-deflection behaviour.
In other embodiments, the biasing arrangement may comprise a plurality of resilient biasing elements, such as springs, configured to apply a desired net resilient biasing force to the pumping element. For example, the biasing arrangement may include two springs configured to counteract one another in such a way that the net resilient biasing force that acts on the pumping element is substantially constant as a function of the position of the pumping element.
The return mechanism may be arranged to apply a force to the pumping element to drive the return stroke of the pumping element. To this end, the return mechanism may be coupled to the pumping element.
The return mechanism may comprise a pivotally-mounted rocker arm driven by a cam. The rocker arm may be arranged to compress the biasing arrangement to remove the resilient biasing force from the pumping element during the filling stroke.
In another arrangement, the fuel pump comprises a driveshaft which is rotatable about an axis, a thrust sleeve slidably mounted for axial movement on the driveshaft, and a connecting member for transferring axial movement of the thrust collar to the pumping element. The biasing arrangement preferably applies the resilient biasing force to the thrust sleeve in the direction of the axis. Conveniently, when the biasing arrangement comprises a pressurising spring, such as a diaphragm spring or a compression spring, the pressurising spring may be mounted concentrically with respect to the driveshaft.
In this arrangement, the return mechanism may comprise a face cam mounted on the driveshaft. The face cam may be arranged to move the thrust sleeve against the resilient biasing force during the filling stroke of the pumping element. In this way, the return mechanism counters the force of the biasing arrangement to allow the filling stroke to occur.
The biasing arrangement may comprise a primary resilient element for applying a primary resilient biasing force to the pumping element to drive the pumping stroke, a secondary resilient element for applying a secondary resilient biasing force to the pumping element to drive the pumping stroke, and adjustment means for adjusting the primary resilient biasing force applied to the pumping element by the primary resilient element. With this arrangement, the pumping stroke of the pumping element can be driven by the secondary resilient biasing force in a first operating condition of the engine, for example at idle speed, and by the primary resilient biasing force applied by the primary resilient element in a second operating condition of the engine, for example at a rated engine speed. The primary resilient biasing force applied by the primary resilient element may be greater than the secondary resilient biasing force, so that the fuel is delivered to the injectors at a higher pressure in the second operating condition than in the first operating condition.
The primary and secondary resilient elements may be springs, such as compression springs, which may conveniently be arranged concentrically. In one example, the primary resilient element is a conical compression spring, and the secondary resilient element is a helical compression spring.
In one embodiment, the adjustment means comprises a seat or carriage upon which the primary resilient element is mounted, and the seat is movable to reduce the primary resilient biasing force applied to the pumping element by the primary resilient element. For example, the seat may be supported on one or more eccentric cams mounted on a shaft which is rotatable to adjust the position of the seat. The shaft may be associated with a throttle of the engine.
The fuel injection system of the first aspect of the invention may include a fuel pump according to the second aspect of the invention. Preferred and/or optional features of each aspect of the invention may be used, alone or in appropriate combination, in the other aspect of the invention also.

Brief description of the drawings

Figure 1 of the accompanying drawings, which is a schematic illustration of a known fuel injection system, has already been described above.
Embodiments of the present invention will now be described, by way of example only, with reference to the remaining accompanying drawings, in which like reference numerals are used for like features, and in which:

Figure 2 is a schematic illustration of a fuel injection system according to an embodiment of the present invention;
Figure 3 is a schematic cross-sectional view of part of a fuel pump suitable for use in the present invention;
Figure 4 is a schematic cross-sectional view of a drive mechanism for a fuel pump according to an embodiment of the present invention;
Figure 5 is a schematic cross-sectional view of a drive mechanism for a fuel pump according to another embodiment of the present invention;
Figure 6 is a plot showing, schematically, the force-deflection behaviour of first and second springs of a drive mechanism for a fuel pump;
Figure 7 is a schematic cross-sectional view of a drive mechanism for a fuel pump according to a further embodiment of the present invention; and
Figure 8 is a plot showing, schematically, the variation of fuel pressure with crank angle for a fuel pump according to another embodiment of the present invention.

Throughout this specification, terms such as 'upper', 'lower', 'upward' and 'downward' are used with reference to the orientation of the parts as illustrated in the accompanying drawings. It will be appreciated, however, that the parts could be disposed in any suitable orientation, in use.

Detailed description of embodiments of the invention

Figure 2 shows a fuel injection system 100 according to an embodiment of the present invention, for use in a compression-ignition internal combustion engine.
The fuel injection system 100 comprises a plurality of electronically-controlled fuel injectors 102 (in this example, two fuel injectors 102 are shown). The fuel injectors 102 are both connected to a single high-pressure fuel pump 104 by way of an injector supply line 106.
The high-pressure fuel pump 104 receives fuel at a relatively low pressure from a fuel tank 110. An electrically-operated transfer pump 112 is provided to transfer fuel from the fuel tank 110 to the fuel pump 104. The transfer pump 112 draws fuel from the tank 110 through a pick-up filter 114, and feeds the high-pressure fuel pump 104 through a fuel feed line 116 which includes an in-line fuel filter 118.
The transfer pump 112 is configured to operate whenever the engine is operating, so that low-pressure fuel is always available in the fuel feed line 116 for delivery to the high-pressure fuel pump 104. A pressure relief valve 120 is connected to the fuel feed line 116 between the transfer pump 112 and high-pressure fuel pump 104 to feed excess fuel back to the tank 110. The outlet of the pressure relief valve 120 is connected to a venturi 122, which provides a low-pressure drain 124 to which the injectors 102 are connected.
The timing of fuel injection from each of the injectors 102 is controlled by an electronic control unit 130. The electronic control unit 130 receives signals from a plurality of sensors that allow the electronic control unit 130 to determine the appropriate injection duration for a given engine operating condition. In this example, the sensors comprise a throttle position sensor 132, an engine speed sensor 134, an engine phase sensor 136, and a temperature sensor 138.
Each injector 102 is of a type generally known in the art. Examples of electronically-controlled fuel injectors that are suitable for use in the present invention include those described in the Applicant's granted European Patent Nos. EP 0 740 068 B , EP 0 798 459 B and EP 1 541 860 B .
In this example, each injector 102 includes a control valve (not shown) which controls the movement of a valve needle (not shown) with respect to a valve seat at the tip of the injector 102. The control valve is operable to increase or decrease the fuel pressure in a control chamber (not shown) disposed at the end of the valve needle opposite the tip, by controlling flow between the control chamber and the low-pressure drain 124. Other electronically-controlled injectors, such as those with directly-acting piezoelectric or solenoid actuators, could also be used.
The high-pressure fuel pump 104 is shown schematically in Figure 3. The pump 104 includes a pumping element or plunger 140 which is slidably received in a bore 142 formed in a pump housing 144. A pump chamber 146 is defined at an upper end of the bore 142, so that the pump chamber 146 is defined, in part, by the upper end of the plunger 140. An inlet passage 148 extends radially though the pump housing 144 to open into the bore 142. The inlet passage 148 receives fuel at relatively low pressure from the fuel feed line 116. An outlet passage 150 extends axially through the pump housing 144 to connect with the pump chamber 146. The outlet passage 150 communicates with the injector supply line 106 by way of an outlet check valve 152.
The pumping plunger 140 is driven by a drive mechanism 160 (not shown in Figure 3). The drive mechanism 160 drives the plunger 140 in linear reciprocal movement with respect to the pump housing 144. In a forward or pumping stroke of the plunger 140, the plunger 140 moves to decrease the volume of the pump chamber 146, and to occlude the inlet passage 148. As a result, fuel in the pump chamber 146 is forced through the outlet passage 150 and the outlet check valve 152 into the injector supply line 106.
At the end of the pumping stroke, the direction of movement of the plunger 140 reverses, and the plunger 140 begins a return or filling stroke. During the filling stroke, the volume of the pump chamber 146 increases and the outlet valve 152 closes. A negative pressure therefore arises in the pump chamber 146 so that, when the inlet passage 148 is uncovered by the plunger 140, fuel flows into the pump chamber 146 from the fuel feed line 116. Operation of the pump 104 continues in a cyclical manner with repeated pumping and filling strokes of the plunger 140.
It is a feature of this embodiment of the invention that the drive mechanism 160 for the high-pressure fuel pump 104 comprises a biasing arrangement, such as a spring, that applies a resilient biasing force to drive the pumping stroke of the plunger 140, and an engine-driven return mechanism for driving the return stroke of the plunger 140. This is in contrast to prior art arrangements, in which the pumping stroke of the plunger is driven by the engine, and the return stroke is driven by a spring.
Figure 4 shows an example of a drive mechanism 160 suitable for driving the high-pressure fuel pump 104 of the system of Figure 2.
The drive mechanism 160 comprises a casing 200 to which the high-pressure fuel pump 104 is mounted. Typically, the drive mechanism 160 would be housed within or partially within the crankcase of the engine. The pump 104 is secured to the casing 200 by a mounting bracket 202 in such a way that a lower end of the pump 104 extends through an aperture 204 in the casing 200.
Within the casing 200, a cup-shaped cap 205 is fitted over the lower end of the plunger 140 of the pump 104. The cap 205 includes a circular base 206 having a recess (not shown) in which the lower end of the plunger 140 is received. A generally cylindrical sleeve 208 extends upwards from the base 206 and slidably engages with a lower portion of the pump housing 144. The base 206 has a larger diameter than the sleeve 208, so that the base 206 forms an outwardly-directed flange of the cap 205. A wire ring or spring clip 210 is provided to retain the base 206 on the end of the sleeve 208. In this way, the cap 205 is guided in linear movement by the pump housing 144, which helps to minimise undesirable side-loading on the plunger 140.
A biasing arrangement in the form of a pressurising spring 220 is disposed between the cap 205 and a spring seat 222 formed on an interior wall of the casing 200. The upper end of the spring 220 bears on the lower side of the base 206 of the cap 205 to apply an upward force to the plunger 140. The lower end of the spring 220 is located in position at the spring seat 222 by a raised land 224. Similarly, the upper end of the spring 220 is located in position on the base 206 by a pin 225. The land 224 and the pin 225 help to guard against lateral movement of the spring 220.
The return mechanism for the plunger 140 comprises a pump driveshaft 230 that extends into the casing 200, and a lever arm or rocker arm 232. The driveshaft 230 is driven by the engine for rotation about an axis that is perpendicular to the axis of movement of the pump plunger 140. The driveshaft 230 carries a cam 234 which, in this example, has three cam lobes. A first end 236 of the rocker arm 232 bears on the cam 234. A second end of the rocker arm 236 is formed into a Y-shape yoke, such that the second end of the rocker arm 236 comprises two fingers 238 (only one of which is visible in Figure 4) that embrace the sleeve 208 of the cap 205 therebetween.
Each finger 238 of the rocker arm 236 has a rounded end 240 that bears on the upper face of the base 206 of the cap 205. The spring 220 biases the base 206 into engagement with the rounded ends 240 of the fingers 238.
The rocker arm 236 comprises a socket 242 that receives a part-spherical pivot member 244. The pivot member 244 is formed at the end of a pin 246 that extends through an aperture 248 in the wall of the casing 200. The pin 246 is held in place by a captive nut 250 that is attached to the outside of the casing 200. Rotation of the pin 246 with respect to the nut 250 allows adjustment of the position of the pivot member 244 within the casing 200.
The socket 242 is disposed intermediate the first end 236 of the rocker arm 232 and the fingers 238 at the second end of the rocker arm 232. In this way, the rocker arm 232 pivots back and forth about the pivot member 244 as the cam 234 rotates.
The fuel pump 160 operates when the cam 234 is driven to rotate in a clockwise direction (as viewed in Figure 4). As each lobe of the cam 234 reaches the first end 236 of the rocker arm 232, the rocker arm 232 pivots in an anticlockwise direction, so that the fingers 238 at the second end of the rocker arm 232 compress the spring 220. In this way, the plunger 140 is retracted downwards to perform the filling stroke of the pumping cycle. In Figure 4, the pump 104 is shown with the plunger 140 in its fully retracted position, in which the volume of the pump chamber 146 (not shown in Figure 4) is maximised and the first end 236 of the rocker arm 232 rests on the nose of the cam lobe.
As the lobe of the cam 234 moves past the first end 236 of the rocker arm 232, the rocker arm 232 pivots in a clockwise direction, under the force of the spring 220. The spring 220 therefore urges the plunger 140 upwards, to perform the pumping stroke of the pumping cycle.
Upward movement of the plunger 140 may be limited by the rocker arm 232 when the first end 236 of the rocker arm 232 rests on the base circle of the cam 234. However, to reduce wear on the rocker arm 232 and the cam 234, it is preferred if the cap 205 is arranged to come into contact with part of the pump housing 144 to limit the upward movement of the plunger 140. In this way, the rocker arm 232 and the cam 234 are separated by a clearance gap when the plunger 140 is at the end of its pumping stroke.
As a consequence of using a spring 220 to drive the pumping stroke of the plunger 140, the output pressure of the pump 160 is determined by the force applied to the plunger 140 by the spring 220.
Accordingly, the spring 220 is designed to apply a substantially constant force to the plunger 140 over the whole range of movement of the plunger 140. In other words, the force-deflection curve for the spring 220 has a relatively flat profile over the operating range of the spring 220. Such constant-force springs of this type are generally known in the art. In the example of Figure 4, the spring 220 is a concially-shaped compression spring, in which the pitch of the coils decreases moving from the lower, largest-diameter end of the spring to the upper, smallest-diameter end of the spring (the change in pitch of the coils is not illustrated in Figure 4).
Referring back to Figure 2, by using a constant-force spring 220, the output pressure of the pump 160, and hence the pressure of fuel in the injector supply line 106 and supplied to the injectors 102, is maintained at a substantially constant, and known, value. Accordingly, neither a rail pressure sensor, nor electronically-controlled metering valves to control fuel flow from the fuel feed line 116 into the pump 104, or from the pump 104 to the injector supply line 106, are required. The fuel injection system 100 is therefore considerably simpler than the known electronically-controlled injection system shown in Figure 1.
Because the pressure of fuel received by the injectors 102 is known, an accurate quantity of fuel can be injected on demand by each fuel injector 102 by controlling the duration of each fuel injection event. Thus the present invention allows fuel injection to be controlled using a pressure-time metering strategy, as is known from more complex systems in the prior art.
In use of the fuel injection system 100, the electronic control unit 130 monitors the engine conditions using the sensors 132, 134, 136, 138 and calculates the optimum volume of fuel for the next injection event for each injector 102. The constant value of the fuel pressure in the injector supply line 106, determined by the force of the spring 220, is preprogrammed into the electronic control unit 130, so that the injection duration required to deliver the desired fuel volume can be calculated.
To start fuel injection from an injector 102, the electronic control unit 130 applies an appropriate control signal to the injector 102 to cause the injector valve needle to open, allowing the high-pressure fuel to be sprayed into the combustion chamber associated with that injector 102. Once the required injection duration has elapsed, the electronic control unit 130 applies another appropriate control signal to the injector 102 to cause the injector valve needle to close, terminating the injection.
The force applied to the plunger 140 by the spring 220 may be adjusted to a required value during assembly of the pump 160 by inserting shims (not shown) between the spring 220 and the spring seat 222, and/or between the spring 220 and the cap 205. The stroke of the plunger 140 can be adjusted by changing the position of the pivot member 244.
The fuel injection system 100 of Figure 2 is particularly suited to small, relatively low power nonroad or utility engines that typically run only at a single rated speed. The spring 220 can be selected to generate an appropriate injection pressure for the rated speed of the engine. For example, to achieve a supply pressure in the injector supply line 106 of 1000 bar, using a plunger 140 with a diameter of 3.5 mm, the spring force selected should be approximately 960 N.
It will be appreciated that the pressure of fuel delivered to the injectors 102 in the system 100 of Figure 2 will not be as precisely defined nor as closely controlled as is the case when a sophisticated electronic pressure control system is used, such as in the known injection system of Figure 1. However, for small utility engines, the pressure control afforded by the present invention is sufficient to realise a substantial reduction in harmful engine emissions compared with mechanical pump-line-nozzle systems.
The single-spring drive mechanism 160 shown in Figure 4 can provide good emissions control for engines which are operated at the single rated speed during almost all of the time that the engine is operating. However, in applications where the engine is run either at a relatively fast rated speed or at a relatively low idle speed, emissions control can be further refined by using a drive mechanism with a two-spring biasing arrangement, as will now be explained with reference to Figure 5.
Figure 5 shows part of a drive mechanism 260 which is similar to the drive mechanism of Figure 4. The fuel pump, plunger, cap and return mechanism have been omitted from Figure 5 for clarity.
The drive mechanism 260 includes a biasing arrangement consisting of a primary conical constant-force spring 320 and a secondary spring in the form of a conventional helical compression spring 360. The constant-force spring 320 is arranged concentrically around the conventional spring 360, hereafter referred to as the inner spring 360, and the upper ends of both springs 320, 360 bear upon the cap (not shown in Figure 5) that receives the plunger of the fuel pump. The constant-force spring 320 is stronger than the conventional spring 360.
A throttle shaft 362 extends through the casing 300 of the drive mechanism 260. The throttle shaft 362 is rotatably mounted in a bearing support 364 housed within the casing 300. The bearing support 364 is fixed with respect to the casing 300, and the lower end of the inner spring 360 bears upon the bearing support 364. In this way, the inner spring 360 always applies a force to the plunger of the pump.
The lower end of the constant-force spring 320 is seated on an annular carriage plate 366, which is movable up and down within the casing 300 in a direction parallel to the direction of movement of the pump plunger. To this end, the carriage plate 366 bears upon a pair of eccentric cams 368 mounted on the throttle shaft 362.
The throttle shaft 362 can be turned from an idle position to a full-speed position using a throttle handle 370. When the throttle shaft 362 is turned, the cams 368 with respect to the carriage plate 366. Figure 5 shows the arrangement with the throttle shaft 362 in its full-speed position, in which the cams 368 are oriented to hold the carriage plate 366 at its uppermost position. In this configuration, the constant-force spring 320 applies a force to the pump plunger. When the throttle shaft 362 is turned to its idle position, the carriage plate 366 moves downwards. This removes the force applied to the plunger by the constant-force spring 320.
In this way, the force applied to the plunger by the drive mechanism 260 can be varied according to the throttle position. At idle speed, only the inner spring 360 acts to drive the pumping stroke of the plunger, such that the pump produces a relatively low outlet pressure (for example of around 250 bar). At the rated engine speed, both the inner spring 360 and the constant-force spring 320 combine to apply a force to drive the pumping stroke of the plunger, resulting in a higher outlet pressure (for example of around 1000 bar).
In this embodiment, the electronic control unit of the engine stores the injector supply pressure for both throttle positions, and switches between the two values according to the throttle position, as determined by the throttle position sensor. In this way, the performance of the injection system can be optimised to minimise emissions at both idle speed and at full rated speed.
The drive mechanisms illustrated in Figures 4 and 5 use conical compression-type constant-force springs to apply a resilient biasing force to the plunger. However, other types of spring with relatively flat load-deflection behaviour can be used. For example, a Belleville washer stack with an appropriate configuration could be used in place of the conical spring.
In another example, a diaphragm spring with substantially constant-force spring behaviour can be used to drive the pumping stroke of the plunger.
It will be appreciated that many constant-force spring designs do not exhibit perfectly flat load-deflection behaviour. For example, some types of spring, such as diaphragm springs, may have a negative load-deflection coefficient over part of the operating range, depending on the spring geometry.
In a variant (not illustrated) of the two-spring drive mechanism shown in Figure 5, a diaphragm spring with a negative load-deflection coefficient over part of the operating range is used in place of the conical spring. At idle, the pumping stroke of the plunger is driven solely by the inner spring, which has linear load-deflection behaviour. At the faster rated engine speed, the diaphragm spring and the inner spring together drive the pumping stroke of the plunger.
Figure 6 shows the force-deflection behaviour of the two springs in this variant. The linear curve of the inner spring is labelled "I", and the non-linear curve of the diaphragm spring is labelled "D". Because the force applied to the plunger by the inner spring is superposed with the force applied to the plunger by the diaphragm spring, the force applied by the inner spring helps to compensate for the negative coefficient region of the force-deflection curve D of the diaphragm spring over the deflection range S corresponding to the plunger stroke.
Figure 7 illustrates another drive mechanism 360 for a fuel pump that could be used in the present invention. The drive mechanism 360 includes a casing 400 that is mounted to an end face 502 of the crankcase 500 of the engine. An output shaft 504 of the engine extends from the end face 502 and through the casing 400, and exits the casing 400 through a shaft seal 460.
The output shaft 504 is driven by the engine crankshaft 506, and acts as a driveshaft for the drive mechanism 360. The output shaft 504 may be integrally formed with the crankshaft 506. The crankshaft 506 is driven by one or more engine pistons 508 (only one piston 508 is shown). Each piston 508 is connected to a respective crank pin (not visible in Figure 7) of the crankshaft 506 by a connecting rod 510.
A face cam 462 is affixed to the output shaft 504 within the casing 400. The face cam 462 comprises an annular collar that rotates with the output shaft 504. One face of the collar is profiled to act as a cam surface for a roller bearing 464, which is located between the face cam 462 and an annular thrust sleeve 466. The thrust sleeve 466 is slidable with respect to the output shaft 504.
In this embodiment, the biasing arrangement comprises a diaphragm spring 468 that is arranged to bias the thrust sleeve 466 into contact with the roller bearing 464 and, in turn, to bias the roller bearing into contact with the face cam 462. The periphery of the central aperture 470 of the diaphragm spring 468 is engaged with the thrust sleeve 466, and the perimeter of the diaphragm spring 468 is located in a recess 512 formed in the end face 502 of the crankcase 500.
As in the embodiment of Figure 4, in the embodiment shown in Figure 7 the high-pressure pump 104 is mounted to the casing 400 of the drive mechanism 360. A connecting member in the form of a bell-cranked yoke 472 is provided to translate movement of the thrust sleeve 466 along the outlet shaft to reciprocal movement of the pump plunger. The yoke 472 comprises a 'U' shaped lower portion with two arms 474 that engage with the thrust sleeve 466, and an angled upper portion 478 that extends at an angle with respect to the arms 474 to engage with a cap 205 provided on the lower end of the plunger of the pump 104. The upper portion 478 of the yoke 478 is coupled to the cap 205 by means of a clip (not shown) or other suitable device. The yoke 472 is pivotally mounted on a pivot shaft 476 that is installed in the casing 400, so that the yoke 472 can pivot back and forth in response to movement of the thrust sleeve 466.
Figure 7 shows the configuration of the drive mechanism 360 at the end of a pumping stroke. As the output shaft 504 rotates from the position shown in Figure 7, the profiled face cam 462 causes the thrust sleeve 466 to be pushed towards the end face 502 of the crankcase 500, against the force of the diaphragm spring 468. This causes the yoke 472 to turn in a clockwise direction, driving the pump plunger in its return stroke.
On continued rotation of the output shaft 504, the spring 468 pushes the thrust sleeve 466 back towards the face cam 462, causing the yoke 472 to turn in an anticlockwise direction, thereby driving the pumping stroke of the plunger using the force of the spring 468.
In the arrangement of Figure 7, the face cam 462 is shaped so that the pump plunger undergoes two pump events and two filling events per revolution of the crankshaft.
In variants of the arrangement of Figure 7, the diaphragm spring can be replaced by a conical compression spring, a Belleville washer stack spring or a similar constant-force spring. In each case, the spring is arranged to apply a spring force to the thrust collar 466 that acts to urge the thrust collar 466 towards the face cam 462.
In the above-described embodiments of the invention, the high-pressure pump delivers fuel to the injectors through injector supply lines. It will be appreciated that the high-pressure pump could instead deliver fuel to a common fuel rail, from which each injector could be supplied. In such an arrangement, the fuel rail acts as an accumulator volume for high-pressure fuel.
In general, the drive mechanism for the fuel pump in the present invention may be configured to provide multiple pumping cycles per revolution of the engine crankshaft. Such a configuration allows rapid re-pressurisation of the fuel in the injector supply lines or, when present, in the common rail.
By way of example, Figure 8 illustrates the variation of fuel pressure in a common rail fed by a fuel pump in a four-stroke engine according to the present invention. A pilot injection (P) is followed by a main injection (M), during which the rail pressure drops considerably. The rail pressure is then restored to its original level in first, second and third pumping events (V1, V2, V3). In this way, the fuel pressure in the rail is maintained at a relatively constant value, and large fluctuations in pressure, which could be detrimental to the fatigue life of the components, are avoided.
It will be appreciated that other simple mechanical biasing arrangements capable of applying a resilient biasing force to the plunger could be employed in place of the spring arrangements described above. For example, any suitable resilient biasing elements, such as resilient rubber bushes, could be used in place of or in addition to the springs. It is also conceivable that the biasing arrangement could comprise a plurality of springs or other resilient biasing elements arranged to cooperate with one another to provide the desired net biasing force on the plunger. For instance, two springs may be arranged to counteract one another so as to provide a net biasing force on the plunger which is substantially constant as a function of the displacement of the plunger.
Many variations and modifications to the invention not explicitly described above can also be contemplated without departing from the scope of the invention as defined in the appended claims.

Claims

A fuel injection system (100) for an internal combustion engine, comprising:
at least one electronically-controlled fuel injector (102);

a fuel pump (104) for supplying pressurised fuel to the or each injector (102); and

a controller (130) arranged to control the injection of fuel from the or each injector (102);

wherein the fuel pump (104) comprises a pump chamber (146), a pumping element (140), and a drive mechanism (160; 260; 360) for moving the pumping element (140) in a pumping stroke in which the volume of the pump chamber (146) is reduced, and a filling stroke in which the volume of the pump chamber (146) is increased and in which fuel is admitted to the pump chamber (146) from a fuel source (116);

characterised in that the drive mechanism (160; 260; 360) comprises a biasing arrangement (220; 320; 468) arranged to apply a resilient biasing force to the pumping element (140) to drive the pumping stroke.
A fuel injection system according to Claim 1, wherein the pressure of fuel supplied to the or each injector (102) is determined by the resilient biasing force applied to the pumping element (140) by the biasing arrangement (220; 320; 468).
A fuel injection system according to Claim 2, wherein the controller (130) is configured to store a pre-determined fuel supply pressure based on the resilient biasing force, and to calculate an injection duration based on the pre-determined fuel supply pressure.
A fuel injection system according to any preceding claim, wherein the resilient biasing force applied to the pumping element (140) by the biasing arrangement (220; 320; 468) is non-linear as a function of deflection of the biasing arrangement (220; 320; 468) over the range of movement of the pumping element (140).
A fuel injection system according to Claim 4, wherein the resilient biasing force applied to the pumping element (140) by the biasing arrangement (220; 320; 468) is substantially constant over the range of movement of the pumping element (140).
A fuel pump for a fuel injection system, comprising:
a pump chamber (146) having an inlet (148) for receiving fuel at relatively low pressure from a fuel source (116), and an outlet (150) for delivering pressurised fuel to the fuel injection system;

a pumping element (140) arranged for reciprocal movement with respect to the pump chamber (146); and

a drive mechanism (160; 260; 360) arranged to drive the pumping element (140) in a pumping cycle comprising a pumping stroke in which the volume of the pump chamber (146) is reduced, and a filling stroke in which the volume of the pump chamber (146) is increased and in which fuel is admitted to the pump chamber (146) from the fuel source (116);

wherein the drive mechanism comprises a biasing arrangement (220; 320; 468) arranged to apply a resilient biasing force to the pumping element (140) to drive the pumping stroke, and an engine-driven return mechanism arranged to counteract the resilient biasing force, thereby to enable the pumping element (140) to undergo the filling stroke.
A fuel pump according to Claim 6, wherein the resilient biasing force applied to the pumping element (140) by the biasing arrangement (220; 320; 468) is non-linear as a function of deflection of the biasing arrangement (220; 320; 468) over the range of movement of the pumping element (140).
A fuel pump according to Claim 7, wherein the resilient biasing force applied to the pumping element (140) by the biasing arrangement (220; 320; 468) is substantially constant over the range of movement of the pumping element (140).
A fuel pump according to any of Claims 6 to 8, wherein the biasing arrangement comprises a compression spring (220; 320) or a diaphragm spring (468).
A fuel pump according to any of Claims 6 to 9, wherein the return mechanism comprises a pivotally-mounted rocker arm (232) driven by a cam (234), and wherein the rocker arm (232) is arranged to compress the biasing arrangement (220; 320) to remove the resilient biasing force from the pumping element (140) during the filling stroke.
A fuel pump according to any of Claims 6 to 9, further comprising a driveshaft (504) which is rotatable about an axis, a thrust sleeve (466) slidably mounted for axial movement on the driveshaft (504), and a connecting member (472) for transferring axial movement of the thrust collar (466) to the pumping element (140), wherein the biasing arrangement (468) applies the resilient biasing to the thrust sleeve (466) in the direction of the axis.
A fuel pump according to Claim 11, wherein the return mechanism comprises a face cam 462 mounted on the driveshaft (504) and arranged to move the thrust sleeve (466) against the resilient biasing force during the filling stroke of the pumping element (140).
A fuel pump according to any of Claims 6 to 12, wherein the biasing arrangement comprises a primary resilient element (320) for applying a primary resilient biasing force to the pumping element to drive the pumping stroke, a secondary resilient element (360) for applying a secondary resilient biasing force to the pumping element (140) to drive the pumping stroke, and adjustment means (362, 366, 368) for adjusting the primary resilient biasing force applied to the pumping element (140) by the primary resilient element (320).
A fuel pump according to Claim 13, wherein the adjustment means comprises a seat (366) upon which the primary resilient element (320) is mounted, wherein the seat (366) is movable to reduce the primary resilient biasing force applied to the pumping element (140) by the primary resilient element (320).
A fuel injection system according to any of Claims 1 to 5, wherein the fuel pump is in accordance with any of Claims 6 to 14.