Abstract:
A two stage pressure limiting valve comprises a valve member arranged for axial movement in a bore. The valve member is biased to close a side spill port and a valve opening communicating with a source of high pressure. Pressure at the valve member/valve seat interface in excess of a threshold value forces the valve member away from the seat whereby a pressure relief volume of fluid is permitted to flow through the valve member itself. Sustained high pressure forces the valve member further away from the valve seat to open a side spill port and establish a larger diversion of fluid at a stable lower pressure level.

Description:
BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention relates to fuel pumps and, more particularly, to fuel pumps and common rail systems for supplying fuel at high pressure for injection into an internal combustion engine. 
     2. Description of the Related Art 
     Modern gasoline fueled automotive internal combustion engines utilize a gasoline direct injection (GDI) system in which highly pressurized fuel, is injected through nozzles directly into each engine cylinder. In a typical GDI system, a high-pressure (200 bar and higher) supply pump is employed which pressurizes fuel received from a low-pressure circuit (2-4 bar) including, e.g., a fuel tank and a low-pressure fuel pump. One such high-pressure supply pump is described in U.S. patent application Ser. No. 09/342,566 filed Jun. 29, 1999, and assigned to the assignee of the present invention. The goal of a GDI system is to inject a vaporized, accurately metered quantity of fuel that is accurately timed for clean combustion. Accurate regulation of the pressure generated by the high pressure supply pump is essential because variations of the supply pressure to the fuel injectors will directly affect both the quantity of fuel and the quality of atomization provided during any given injection event. U.S. patent Ser. No. 09/638,286 filed Aug. 14, 2000 describes a self-regulating gasoline direct injection system in which pressure detection and feedback systems are used to stabilize the supply pressure for a common rail fuel injection system. The self-regulating system monitors pressure in an accumulator for the common rail, adding pressurized fuel when needed and diverting the output of the high-pressure supply pump at a lower pressure when pressure in the accumulator is adequate. This system avoids wasteful pressurization of fuel when it is not needed, saving energy and avoiding excessive heat generated by the depressurization of unnecessarily pressurized fuel. 
     It is known that forced re-circulation of highly pressurized fuel into a high pressure supply pump for a GDI system will quickly overheat the GDI pump and possibly result in catastrophic failure. Therefore, extended periods of forced high-pressure re-circulation must be avoided. In addition, failure of the primary pressure regulator or some other GDI component can result in pressures in the GDI system exceeding the design objectives of components resulting in leakage and/or failure. 
     Thus, there is a need in the art for a pressure limiting valve for a GDI pump that is responsive to excessive pressure having a duration that indicates system malfunction. 
     SUMMARY OF INVENTION 
     An object of the present invention is to provide a new and improved two-stage pressure limiting valve for a GDI pump that prevents pressure related failure of GDI components. 
     Another object of the present invention is to provide a new and improved pressure limiting valve for a GDI pump which absorbs short duration pressure spikes without affecting overall GDI system performance. 
     A further object of the present invention is to provide a new and improved two stage-pressure limiting valve for GDI pump capable of diverting the large flow of pressurized fuel resulting from failure of a primary pressure regulator or other GDI system component. 
     These and other objects of the invention are achieved by a two-stage pressure limiting valve in accordance with the present invention. A preferred embodiment of the two-stage pressure limiting valve comprises a cup-like plunger with an integrated hemispherical ball check member positioned adjacent a complementary valve seat. The plunger is arranged for reciprocal movement in a bore defined by the pump housing. The plunger forms a barrier between a first hydraulic chamber surrounding the ball check and valve seat (the valve chamber) and a second hydraulic chamber within and beneath the plunger. The ball check end of the plunger defines a narrow gage fuel flow passage connecting the valve chamber to the interior of the plunger. A control spring disposed in the plunger bore biases the plunger and its associated ball check against the valve seat. The valve seat defines an opening which is exposed to the high-pressure output passage of a supply pump. A further hydraulic passage communicates between the plunger bore and the interior of the pump housing, i.e., the sump. 
     The plunger, plunger bore and hydraulic passage to the sump are configured to provide two alternative fluid flow paths. A first, limited volume path is defined through the narrow gage opening in the plunger and around or through the plunger skirt to the sump passage. This first path does not require significant displacement of the plunger within its bore. A second, large volume path is opened when the plunger is forced back in its bore against the force of the control spring. When the plunger moves away from the valve seat a pre-determined distance, the outer periphery of the plunger acts as a valve to uncover the sump passage. The second, large volume path extends directly from the valve chamber into the sump passage. 
     Under normal engine operating conditions, e.g., when fuel pressure at the output passage of the supply pump is below a pre-established upper limit, the ball check will remain firmly seated against the valve seat by the bias spring. In the event of a short duration pressure spike, the ball check will lift from its seat and a small quantity of fuel to be vented into the valve chamber. The vented fluid will then pass through the narrow gage passage to the interior of the plunger and subsequently into the sump passage. When the output pressure of the supply pump exceeds the pre-established upper limit for an extended duration, the narrow gage passage in the plunger is no longer capable of diverting the volume of fuel necessary to reduce pressure to an acceptable level. The excess fuel accumulates in the valve chamber, forcing the plunger away from the valve seat and opening the second large volume fuel pathway into the sump passage. The plunger will remain in this position to divert the large quantity of fuel necessary until the problem causing the excess pressure is corrected. 
     Collapse of the control spring due to excessive pressure permits the plunger to move to a position where large quantities of fuel are re-circulated into the pump housing. This re-circulation position represents a new stable state at a much reduced pressure, e.g., 30 bar, from the normal operating pressure of the supply pump, e.g., in excess of 200 bar. The GDI electronic control unit (ECU) may be programmed to detect this new lower stable state condition and place the GDI system in a limp home mode, permitting the vehicle to be driven to the closest service station for repair of the underlying problem. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     These and other objects, features and advantages of the invention will become readily apparent to those skilled in the art upon reading the description of the preferred embodiments in conjunction with the accompanying drawings in which: 
     FIG. 1 is a sectional view through a two-stage pressure limiting valve in accordance with the present invention; 
     FIG. 2 shows the two-stage pressure limiting valve of FIG. 1 responding to a pressure spike; 
     FIG. 3 shows the two-stage pressure limiting valve of FIG. 1 responding to a long duration over-pressure condition; and 
     FIG. 4 is a schematic diagram illustrating the two-stage pressure limiting valve in the context of a simplified gasoline direct injection system. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
     With reference to the drawings in which like numerals represent like parts throughout the several Figures, a two-stage pressure limiting valve in accordance with the present invention is generally designated by the numeral  10 . FIG. 4 illustrates the two stage pressure limiting valve  10  in the context of a simplified gasoline direct injection system including a high pressure supply pump  8  and common rail  4 . The high pressure supply pump  8  is provided with low pressure fuel through a feed line  6 . Low pressure fuel is drawn from the sump  12  and pressurized by pumping means  5 . High pressure fuel is fed to the common rail  4  through the high pressure output passage  14  of the pump  8 . Metered quantities of fuel are released from the pressurized common rail  4  into the combustion chambers of an internal combustion engine (not shown) by the injectors  2 . The two stage pressure limiting valve  10  is arranged to limit the pressure delivered to the common rail  4  by diverting fluid back to the sump  12  through a low pressure sump passage  28 . A two-stage pressure limiting valve in accordance with the present invention may be used in association with any high-pressure pump whether or not the pump is equipped with a primary pressure regulator. Therefore, the configuration and operation of the high-pressure supply pump and/or primary pressure regulator will not be further discussed herein. 
     A preferred embodiment of the two-stage pressure limiting valve  10  may be incorporated into the housing  40  of a high-pressure supply pump (as illustrated herein) or may be provided as a separate component. The pump housing  40  defines a sump chamber  12 , which is typically filled with fuel at a relatively low feed pressure of between 2 and 4 bar. The pressurizing mechanism of the pump (not shown) draws low pressure fluid from the sump chamber  12 , pressurizes the fuel to a typical pressure of 200 bar or above, and delivers the pressurized fuel to a high pressure output passage  14 . 
     The illustrated preferred embodiment of a two-stage pressure limiting valve  10  comprises a plunger  18 , a valve seat  16 , and a control spring  20 . The plunger  18  and control spring  20  are arranged in a bore  25  defined by the pump housing  40 . The cup-shaped plunger  18  includes a skirt  19  projecting axially away from the valve seat  16 . The control spring  20  is surrounded by the plunger skirt  19  and is arranged to bias an integral hemispherical ball check  17  against a complementary valve seat  16 . The spring  20  is preferably a constant rate coil spring selected to minimize rail pressure variation during the first stage of valve operation. The fluid passage  22  in the valve seat  16  defines a first “active area” or area of the plunger exposed to rail pressure. This first active area is utilized during the first stage of valve operation. When the volume of fluid passing through the fluid passage  22  exceeds the volume capacity of the narrow gage hydraulic passage  23 , the plunger  18  is forced away from the valve seat to expose a second, larger “active area” exposed to the rail pressure. This second active area comprises the valve end of the plunger  18 . It will be understood that an equivalent rail pressure acting on the larger second active area will produce a correspondingly larger force on the plunger  18 . 
     The valve seat  16  defines a fluid passage  22  in communication with the high-pressure output passage  14  of the pump. A valve chamber  24  is defined at the end of the bore  25  adjacent the valve seat  16 . A narrow gage hydraulic passage  23  through the plunger  18  connects the valve chamber  24  with a second hydraulic chamber  27  defined by the plunger skirt  19  and plunger bore  25 . A sump passage  28  connects the bore  25  with the sump chamber  12 . One portion  26  of the bore  25  has an enlarged diameter, whereby a coaxial hydraulic passage  29  is defined between the piston skirt  19  and the pump housing  40 . The coaxial hydraulic passage  29  permits fluid flow from the second hydraulic chamber  27  into the sump passage  28 . 
     FIG. 1 illustrates the relative positions of the plunger  18 , valve seat  16 , and control spring  20  under normal pump operating conditions. Restated, FIG. 1 illustrates the relative positions of the components of the two-stage pressure limiting valve when the output pressure generated by the pump is below some pre-established maximum, e.g., 200 bar. It should be understood that fuel pressure at the high-pressure output passage  14  of the pump may frequently exceed the pre-established upper limit for brief periods. FIG. 2 illustrates the relative positions of the components of the two-stage pressure limiting valve in response to such a short duration pressure “spike”. 
     The term “spike” as used in this application is defined as a short duration pressure rise, lasting for a small percentage of the duration of one system cycle. The duration of a typical pressure spike will be measured in microseconds, while the system cycles are typically measured in milliseconds. Spikes are caused by sudden events in the hydraulic system, for example, sudden changes in flow velocity, sudden change in flow direction, or the impact of a valve on its seat (creating a hydraulic pressure wave known as a “water hammer”), etc. Spikes created by these events propagate by wave motion travelling at the speed of sound through the entire hydraulic system. Occasionally, pressure waves from different sources (or reflected waves from the same source) can superimpose on one another, resulting in pressure spikes having an effective pressure corresponding to a multiple of the nominal system pressure. Pressure spikes are in contrast to longer lasting pressure rises typically referred to as pressure “surges”. 
     A pressure spike will cause the ball check  17  to lift from its seat  16  and vent a small amount of fuel into the valve chamber  24 . From the valve chamber  24 , the vented fuel passes through the narrow gage hydraulic passage  23  and into the second hydraulic chamber  27 . The vented fuel then flows radially outwardly and axially through coaxial passage  29  as indicated by the dashed line and arrow of FIG.  2 . 
     Thus, FIG. 2 illustrates the first stage of the two-stage pressure limiting valve. During the first stage, small quantities of fuel can be vented from the high pressure output passage  14  of the pump through the valve seat  16 /ball check  17  interface, valve chamber  24 , narrow gage hydraulic passage  23 , second hydraulic chamber  27 , coaxial passage  29  and sump passage  28  to return to the pump sump chamber  12 . When the pressure spike has passed, control spring  20  re-seats the ball check  17  against the valve seat  16  and the GDI system is permitted to continue functioning as normal. 
     In the event of a more significant failure, for example, failure of the primary pressure regulator or some major fuel injection component, pressure at the high-pressure output passage  14  of the pump may exceed the pre-established limit for an extended duration. Under such circumstances, the volume of fuel that must be re-circulated to relieve the overpressure condition will be greater than the amount of fuel that can pass through passage  23  as illustrated in FIG.  2 . FIG. 3 illustrates the relative positions of the valve seat  16  and plunger  18  in response to a pressure surge or overpressure condition of extended duration. The initial surge of pressure will result in relative positions as illustrated in FIG.  2 . However, the volume of fuel entering the valve chamber  24  will exceed the volume of fuel which can pass through the narrow gage hydraulic passage  23 . Therefore, the volume of fluid in chamber  24  will increase, forcing the plunger  18  away from the valve seat  16  and ultimately collapsing the control spring  20 . 
     Movement or displacement of the plunger  18  away from the valve seat  16  causes the upper shoulder of the plunger to open a second, larger fluid passage or side spill port directly from the valve chamber  24  into the sump passage  28 . So long as the volume of fluid entering the valve chamber  24  exceeds the volume of fluid which may pass through the narrow gage hydraulic passage  23 , the relative positions of the plunger  18  and valve seat  16  will remain those illustrated in FIG.  3 . Fluid flow under these circumstances is illustrated by the dashed line and arrow in FIG.  3 . 
     If the conditions that produce excessive pressure at the pump high pressure output passage  14  are substantially permanent, the component positions illustrated in FIG. 3 will be maintained, establishing a new stable state at a pressure level of preferably  25  and 35 bar. As soon as the electronic control module for the GDI system detects this stable reduced pressure level, the ECU will enter a limp home mode where the injection is advanced to permit the affected vehicle to be driven to the nearest service station for repair. When the vehicle is turned off and the excessive flow through the valve seat orifice  22  is stopped the plunger  18  and associated ball check will automatically re-seat and normal GDI operation can resume, assuming that the underlying problem has been corrected. 
     During stage one of valve operation, the pressure is regulated at the valve member  17 /valve seat  16  interface as a balance between the hydraulic force acting over a small exposed plunger area and a pre-determined spring force. During stage two of valve operation, the valve member is far away from the valve seat and pressure regulation occurs as a balance between hydraulic force acting over the larger frontal area of the plunger  18  and a slightly higher spring force exerted by the now compressed control spring  20 . One working example is a high pressure supply pump having a normal output pressure of 200 bar and a two stage pressure limiting valve designed to have a threshold pressure pressure 20 to 30 bar above the normal output pressure of the pump. The threshold pressure may typically be between 10 and 20% above the normal rail operating pressure. 
     The flow volumes triggering the transition between first and second stage valve operation will depend on the nominal output volume and pressure of the high pressure supply pump. Another factor is the maximum heat release (from re-circulated high pressure fuel) that can be tolerated without creating vapor cavities in the sump of the pump and/or without compromising the integrity of pump components. The relationship between the first and second flow volumes may be manipulated by selection of the following parameters: diameter of plunger  18 , flow area across valve seat  16 , flow area of the narrow gage hydraulic passage  23 , spring rate of spring  20  as well as the location and geometry of the sump passage  28 . As an initial design parameter, the transition between first and second stage valve operation may be selected to occur at approximately 10% of the nominal pump output volume at maximum speed. Although the relative percentile of this transition flow volume will increase at lower pump speeds, the total amount of released heat will also decrease. The second flow volume may be between 8 and 10 times the first flow volume. 
     While a preferred embodiment of the invention has been set forth for purposes of illustration, the foregoing description should not be deemed a limitation of the invention herein. Accordingly, various modifications, adaptations and alternatives may occur to one skilled in the art without departing from the spirit and the scope of the present invention.