Pressure relief valve for single plunger fuel pump

A fuel pump including a low pressure infeed leading to a pumping chamber, a pumping plunger for pressurizing fuel in the pumping chamber, an outlet valve for delivering pressurized fuel from the pumping chamber to a high pressure outlet line during pumping, and an overpressure relief valve connected between the high pressure outlet line and the pumping chamber. A ball type relief valve element is spring biased to close against the relief valve seat. The spring is in a spring chamber that is fluidly isolated from the pumping chamber while the valve element is closed, preferably in direct hydraulic communication with the low pressure infeed, thereby reducing dead volume during normal pumping. The spring can act on the valve element through a piston, and in an alternative embodiment, the piston can provide a spill path from the pumping chamber to the low pressure infeed.

BACKGROUND

The present invention relates to high pressure fuel supply pumps for gasoline engines.

Single piston, cam driven fuel pumps have become a common solution for generating high pressure fuel for common rail direct injection gasoline engines.

It is known in the industry that the pump must incorporate an outlet check valve to prevent pressure bleed back from the rail while the pump is in the intake stroke cycle. It has become an industry requirement to incorporate a pressure relief valve within the pump to protect the entire high pressure system from an unexpected excess pressure caused by a system malfunction. In order to protect the rail and fuel injectors, the pressure relief valve must be in hydraulic communication with the rail. Two executions of such hydraulic communication, in parallel with the pump flow, are described in U.S. Pat. No. 7,401,593 and U.S. Pat. No. 8,132,558. The executions described in the prior art are successful in their ability to achieve a reasonable relief pressure by hydraulically disabling the relief device during the pumping event when normal high pressure line pulsations occur.

While these executions are sufficient for current gasoline direct injection systems that operate up to about 200 Bar rail pressure, there is a significant limitation for future systems that will operate at higher pressures required to meet forthcoming emissions regulations. Because the pressure relief valve flow returns to the pumping chamber, its associated spring and spring cavity are in direct communication with the pumping chamber. This spring cavity adds significant dead volume to the pumping chamber circuit volume that must be compressed during each pumping event. Higher operating pressures require increased pressure relief valve opening pressures, higher spring loads, and increased spring cavity volume to accommodate the increased spring size. This added dead volume combined with the increased pumping pressures has a significant detrimental effect on pump efficiency.

SUMMARY

The present invention provides a solution with the same pressure relief valve function, but with a reduced pumping chamber circuit volume, and thus improved efficiency. This is accomplished by isolating the pressure relief valve spring chamber from the pumping chamber, such as by locating the pressure relief valve spring chamber within the low pressure side of the pump.

The spring load to the pressure relief valve is preferably applied through a close-fitting piston wherein the spring side of the piston is exposed to the low pressure side of the pump, and the valve side of the piston is exposed to pumping chamber pressure.

In an embodiment where the piston acts on a ball type relief valve, the piston sealing diameter should be less than or equal to the pressure relief valve sealing diameter in order to achieve a reasonable relief valve opening pressure.

According to a refinement, a secondary relief function is provided to accommodate a severe system malfunction. This added relief function would become enabled when the relief valve piston is advanced to a position associated with very high pumping chamber pressures. This function is achieved by the addition of a drilling through the relief valve piston that becomes uncovered from the piston bore in the advanced position, thereby connecting the pumping chamber with the low pressure side of the pump, relieving the excess pumping chamber pressure.

Since the volume of the relief spring chamber for high pressure systems is much larger than the volume of the related flow passages and relief valve chamber, eliminating the relief spring chamber from the cyclic pressurization of the pumping chamber circuit volume resulting from the reciprocation of the pumping plunger significantly reduces dead volume and thereby increases efficiency.

DETAILED DESCRIPTION

Two representative embodiments will be described with reference to the accompanyingFIGS. 1-7.

FIG. 1is an overall system schematic illustrating the fuel system for an internal combustion engine as described in U.S. Pub. US 2011/0126804. The low-pressure pump2pressurizes fuel from the fuel tank1, and delivers it at low feed pressure to the high pressure pump3through an inlet fitting. The fuel then passes by an accumulator4, and continues at low pressure through passage2′ to a normally open inlet control valve5. A normally closed control valve is also applicable to such a fuel system. The fuel is then drawn into the pumping chamber10, where it is pressurized by the upward motion of the pumping plunger8as reciprocally driven by the engine camshaft9. The inlet control valve5is acted upon by the control valve spring7and solenoid6to control the quantity of fuel delivered by the high pressure pump. This is accomplished by the accurate timing of the control valve closing relative to the pumping piston upward travel position.

The pressurized fuel travels through the outlet check valve11, high pressure line14, and into the common rail16that feeds the engine fuel injectors15. Because the injectors are fed from a common rail, injector timing is flexible. Desired rail pressure is controlled by a closed loop electronic control unit (ECU)18, based on feedback and control of the high pressure fuel output via the solenoid6and control valve5compared to the rail pressure sensor17output signal to the ECU18. A pressure relief valve12is required to protect the high pressure system in case of a system malfunction. The outlet check valve and pressure relief valve are preferably in a common fitting assembly13, but this is not required for the present invention.

FIG. 2is a schematic incorporating a first embodiment of the relief valve in accordance with the present invention. The normal pump operation is the same as the pump described inFIG. 1. In this embodiment however, the relief valve12is urged against its sealing seat via a spring located in the low pressure side of the pump and a relief valve spring isolation piston19, which also separates the pumping chamber pressure10from the low pressure side of the pump. During a system malfunction, the excess pressure in the rail16and line14opens the relief valve12, moving the relief valve spring isolation piston19, and flowing fuel back to pumping chamber10during the charging cycle of the pump. In this embodiment, the pressure relief spring cavity20is in direct fluid communication via passage20′, entirely within the pump, with the inlet line or passage to the inlet control valve5.

FIG. 3is a schematic incorporating a second embodiment of the relief valve in accordance with the present invention. In this embodiment, excessive pressure in the pumping chamber10(caused by a severe high speed system malfunction for example), induces the relief valve spring isolation piston19to retract in its bore to a position allowing the overpressure spill channel19′ to fluidly connect the pumping chamber10with the low-pressure spring chamber20and thereby spill chamber pressure into the low pressure side of the pump.

FIG. 4shows one execution of the pump described with respect toFIG. 2, andFIG. 5shows one execution of the second embodiment of the pressure relief valve described with respect toFIG. 3.

InFIG. 4, a ball type relief valve12seals against the pressure in the high pressure line14via flow channels including flow passage14′. Also included in the execution is a high pressure sealing plug21.FIG. 6is a detailed view in the area of the pressure relief valve ofFIG. 4. It is advantageous to have the sealing diameter D2of the piston19less than or equal to the sealing diameter D1of the ball12against the ball seat. This prevents unwanted motion of the piston19during a normal pumping event when pumping chamber10and channel10′ pressures spike above rail pressure at high speeds, but also allows a reasonable opening pressure of the relief valve12during the pump charging event, or during a hot soak. The spring22can operate directly on the piston19or as shown, through an intermediate spring seat23.

InFIG. 5, the piston19includes an overpressure spill connection19′ defined by19a,19b, and19c. This is shown in closer detail inFIG. 7. The piston19is mounted in the bore of sleeve24. The sleeve has a front cavity10′ with ball valve12in fluid communication via channel10″ with the pumping chamber10and a back cavity26in fluid communication with spring chamber20. The back portion of the piston19extends through back cavity26into spring cavity20for loading by spring22. The piston has a front port19aleading to a central bore19bwhich fluidly connects to a back port19c. The bore19bextends only part way through the piston19, with back port19cblocked by the sleeve ID when the valve12is seated and when valve12lifts off seat28, moving the piston19a first, relatively short distance (less than ‘X’), to provide the pressure relief from high pressure line14,14′ back into the pumping chamber10via passage10″.

During a severe system malfunction, the pressure in pumping chamber10and channel10′ can overcome the closing pressure defined by the load from spring22and the piston diameter, thereby moving the piston19by at least distance ‘X’. This connects pumping chamber pressure in channel10″ and chamber or cavity10′ to the low pressure in back cavity26and spring chamber20, thus spilling pumping chamber pressure from10to the low pressure side of the pump. The piston19will advance past distance ‘X’ during the pumping stroke if the hydraulic force is sufficient to overcome the force from spring22. The piston is thus displaceable in the bore a second distance greater than ‘X’ from seat28by the force of fuel pressure in chamber10acting against piston19during a pumping stroke. During that condition the pressure acts independently of the ball, against the OD and back wall of bore19bto advance the piston until port19cis exposed.

If the rail pressure is too high and relieves into the pumping chamber during the next charging stroke, the pumping chamber10will have higher pressure than it would normally have when beginning the next pumping stroke. Depending on the RPM and total flow being recirculated, it can begin to “back up”, driving the rail pressure to a much higher level. In the embodiment ofFIGS. 2 and 6, the overpressure will dump back to the pumping chamber; this will keep flowing in a closed circuit10,11,14′,12,10′,10″,10. In the embodiments ofFIGS. 3 and 7the pumping stroke overpressure would spill through10′,10″,19,20,20′ and2′. In each case the rail pressure will stabilize for a given RPM and flow condition. The higher the RPM and flow rate, the higher the stabilized the rail pressure. The ball valve will lift during the charging stroke when pressure in14′ exceeds the set pressure, but could reclose during the pumping stroke independent of the piston19, only if the pumping chamber pressure is sufficiently high to move the piston19. In that case the ball is “free floating” and will likely close due to the pressure differential across the ball during the pumping stroke. As stated above, there is a condition where the piston19can separate from the ball12to dump fuel through passage19′. In that case the spring22is not acting against the ball.

FIGS. 6 and 7show a transverse passage10″ between the pumping chamber10and the cavity or chamber10′ for the ball type relief valve12. This passage10″ can be a direct hole through the sidewall of the chamber10′ behind the ball valve12(as depicted schematically inFIG. 2), or the passage10″ can open into the chamber10′ just behind the seat for ball valve12. InFIGS. 6 and 7, the flow path between the pumping chamber10and the cavity or chamber10′ includes at least one fluted channel10″′ extending longitudinally along the outside of the front end of the sleeve24, connecting the pumping chamber10to the transverse passage10″. Since the ball valve12is situated loosely within the sidewall wall of chamber10′, it is not critical where the flow passage10″ enters chamber10′. The ID of the sidewall of chamber10′ can be sized to guide the ball12when it opens. This eliminates the possibility of the ball becoming permanently disengaged from the seat. It should also be appreciated that the spill connection19′ can take other forms.

A key aspect of the present invention is that the spring chamber20is fluidly isolated from fuel pumped in the pumping chamber10while the valve element12is closed. In the embodiment ofFIGS. 2, 4, and 6(wherein no overpressure spill path19′ is provided), spring chamber20can remains isolated from pumping chamber10whether or not the valve element12open. In the embodiment ofFIGS. 3, 5, and 7(wherein a pumping overpressure spill path19′ is provided), spring chamber20can be fluidly connected to the pumping chamber20. Notwithstanding that chamber10′ at the front of the sleeve24is subject to the pumping pressure in chamber10, in the illustrated embodiments, this isolation of the spring chamber20while the valve element12is closed is achieved by sealing diameter D2of the central portion of piston19closely sliding within the central portion of sleeve24(as shown inFIGS. 6 and 7). If the overpressure spill connection19′ is not to be implemented, the back cavity26is not needed and the sleeve24can be further simplified.