Abstract:
A high pressure fuel supply pump comprising a housing ( 102 ) and a pump body ( 108 ) fixed within the housing along a body axis and including a plurality of radially oriented plunger bores ( 20 ), each bore having a plunger ( 22 ) disposed therein. An actuating assembly ( 116 ) is disposed around the plunger bores for producing reciprocal motion. A central cavity ( 120 ) extends along the axis and intersects the pumping bores to form a pumping chamber ( 18 ) in cooperation therewith. An inlet check valve ( 132 ) and a discharged check valve ( 124 ) are in communication with the pumping chamber.

Description:
This application claims the benefit under 35 USC §119 (e) of U.S. Provisional Application 60/076,373 filed Feb. 27, 1998. 
    
    
     BACKGROUND OF THE INVENTION 
     The present invention relates to high pressure hydraulic pumps, and particularly to pumps for supplying diesel fuel at high pressure in a fuel injection system for vehicles. 
     Rotary hydraulic pumps for use in diesel fuel injection systems for internal combustion engines, have been well known for a number of years. Recently, desired improvements in fuel efficiency and emissions control, have led the automotive industry toward development of so-called common rail fuel injection systems, whereby a high pressure pump is utilized to establish and maintain a high fuel pressure in an accumulator in fluid communication with individual injectors. Individual injection events are controlled at the injectors for achieving combustion in the individual combustion chambers of the internal combustion engine. This is in contrast to the more common distributor type fuel injection pumps, whereby fuel pulses are distributed from within the pump, to individual distribution paths leading to a respective plurality of injectors. 
     Common rail pumps are expected to operate at about 20,000 psi, whereas conventional distributor pumps operate at less than about 10,000 psi. This difference accentuates certain drawbacks in conventional pumps, such as an excessive amount of fuel that experiences pressurization in connection with the pumping action, and the excessive amount of heat carried by fuel which pressure pumping, but which is not actually injected into the combustion chambers. 
     Unfortunately, many of the disadvantages of distributor type pumps in this regard, have been carried over into attempts to modify the distributor type pumps, for use in common rail systems. The problem of excess pumping and associated heat generation, arise especially in the so-called pump-spill, spill-pump-spill, and fill-spill techniques, as exemplified in commonly owned U.S. Pat. Nos. 5,215,449 and 5,688,110. The reluctance in setting aside such spill-type pumps, is that the fuel delivery requirements on the pump can vary considerably depending on, for example, whether the pump is starting from a cold condition, whether the pump is running at a sustained, steady state condition, and whether acceleration is required to handle an increased load. With the spill-type pumps, a quantity of fuel is delivered to the pump in an amount greater than any necessary requirement, and spill control is utilized during pumping, to try to match the quantity discharged from the pump, with the instantaneous requirements. 
     Other techniques attempt to match fuel quantities delivered to the pumping chamber, with the instantaneous requirements, e.g., pre-metering based on computations of pump demand by an electronic control unit (ECU). This pre-metering of the fuel quantity to be charged into the pumping chamber is typically controlled by a solenoid valve responsive to a control signal from the ECU. A major disadvantage of solenoid-implemented pre-metering, is the relatively long duration required for the metering of a useful quantity of fuel through the solenoid valve, and the difficulty to adjust the metered quantity over a wide range according to the needs of the engine. In many instances, the intake phase of pumping chamber operation with pre-metering, would not leave sufficient time to implement the pumping phase using a cam pumping rate profile shallow enough to assure quiet operation. Even with dual-rate pumping profiles, there is not enough time available during the pumping phase of a cycle, to incorporate such duality. 
     Another consideration which leads to significant disadvantages in the use of conventional distributor pumps for common rail injection systems, is the relatively long fuel flow paths associated with feed and discharge phases of operation. 
     SUMMARY OF THE INVENTION 
     It is, therefore, an object of the present invention to provide a high-pressure hydraulic pump which minimizes the quantity of fuel charged in the pumping chamber during the intake phase of operation, is highly energy efficient during steady state pumping operation, yet in the preferred embodiment can respond quickly to transients, such as acceleration. 
     It is another object, notwithstanding the nature of the manner in which fuel is fed to the pump, that the flow passages within the pump minimize dead volume and provide for highly efficient charging and discharging valve operation. 
     The objects set forth above are satisfied according to the present invention, by a cam having an internal actuation profile for the simultaneous inward actuation of a plurality of pumping plungers, whereby the fuel is pressurized in a substantially common, central pumping chamber, from which the high pressure fuel is discharged through an outlet located substantially on the drive shaft axis. 
     In one aspect, the invention has a pump housing which includes a substantially cylindrical cavity in which a stationary body portion is mounted, the body including a stationary hub portion which carries the shoe and roller assemblies radially outwardly of the plunger bores. The body includes another longitudinal, central cavity, which contains a high pressure outlet fitting. An axially slidable control valve is supported within the fitting. An inlet chamber and the inlet check valve are situated along the axis of the fitting. The inlet check valve and the discharge check portion of the control valve are both located close to the plane passing through all the pumping plungers. As a result, the flow passage between the inlet check valve and the pumping chamber, and from the pumping chamber to the discharge check valve portion, are relatively short, minimizing dead volume. Furthermore, the substantially axial flow path from the discharge check valve, along the control valve and through the fitting for discharge through a single outlet at over 100 bars, provides significant efficiencies and advantages. 
     The invention may also be considered as a high pressure fuel supply pump, comprising a pump housing and a pump body fixed within the housing along a body axis and including a plurality of radially oriented plunger bores, each bore having a pumping plunger disposed therein for reciprocal radial motion. An actuating assembly is disposed within the housing and around the plunger bores for producing the reciprocal motion by simultaneously driving the plungers radially inwardly during a pumping phase of the operation and simultaneously retracting the plungers radially outwardly during a charging phase of operation. A central cavity extends along the axis and intersects the pumping bores to form a pumping chamber in cooperation therewith. A feed fuel supply train includes an inlet check valve biased to open and fluidly expose the plunger bores to a supply of feed fuel at a relatively low pressure during the charging phase of operation and to seal against the supply of feed fuel during the pumping phase of operation. A high pressure outlet fitting is fixed in the central cavity and includes an internal valve cavity in fluid communication with the pumping chamber and extending along the axis. A discharge check valve is biased to seal the valve cavity from the pumping chamber while the inlet check valve opens to deliver low pressure fuel to the pumping bores during the charging phase of operation and to fluidly expose the valve cavity to the pumping chamber during the pumping phase of operation. In this manner, during the pumping phase of operation, high pressure fuel is discharged from the pumping chamber through the outlet fitting substantially along the axis. 
     The preferred implementation includes the steps of pre-metering successive quantities of fuel from a reservoir to a positive displacement transfer pump, then actuating the transfer pump to raise the pressure of the successive quantities of fuel by at least about 100 psi, preferably 200-300 psi. Each quantity of fuel which was pressurized in the transfer pump, is delivered to a high pressure pumping chamber defined in part by a plurality of fluidly interconnected high pressure pumping bores, so that each pumping bore receives a certain, i.e., predetermined, charge of fuel within a first time interval. A plurality of plungers in the respective pumping bores are then simultaneously actuated to increase the pressure in the pumping chamber to the desired high pressure, preferably at least about 15,000 psi, within a second time interval, and discharge the quantity of fuel through a high pressure discharge valve. The second time interval is of longer duration than the first time interval. The first time interval can be relatively short, because the pumping chamber is charged by the transfer pump at a pressure of at least about 100 psi, which is considerably higher than the conventional charging pressure. As a result, the necessary quantity of fuel can be delivered to the pumping chamber in a relatively short time period. Therefore, each pumping plunger can be actuated by a dual rate cam profile over a relatively long time period such that at steady state the actuation occurs only along a relatively shallow slope of the cam profile, whereas when acceleration is required, the actuation can occur more quickly, along a steeper profile, before continuing along the relatively shallow profile. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     These and other objects and advantages of the invention will be described in greater detail below with reference to the accompanying drawings, in which: 
     FIG. 1 is a schematic of a portion of a common rail fuel injection system incorporating one embodiment of the high pressure pump assembly of the present invention; 
     FIG. 2 is a schematic of the interaction of the transfer pump and high pressure pump for the timing of maximum fuel delivery during steady state operation and constant accumulator pressure; 
     FIG. 3 is a schematic of the interaction of the transfer pump and high pressure pump for the timing of maximum fuel delivery during transient operation and simultaneous accumulator pressure increase; 
     FIG. 4 is a longitudinal section view of a common rail supply pump assembly for implementing the features shown schematically in FIGS. 1-3; 
     FIG. 5 is an illustrative cross section view of the pump shown in FIG. 4; 
     FIG. 6 is a schematic of the charging operation of the components of the pump shown in FIG. 4; 
     FIG. 7 is a schematic of an alternative to the pump shown in FIG. 6; and 
     FIG. 8 is a longitudinal section view along a plane different from that shown in FIG. 7, revealing the feed path to the inlet chamber. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     FIG. 1 shows schematically the essence of the operating principle of a high pressure fuel supply system  10 , according to the present invention. In the illustrated embodiment, the system  10  is arranged to supply high pressure fluid, such as diesel fuel, to an accumulator  12 , for ultimate injection into a diesel engine (not shown). In such a so-called common rail fuel injection system, the accumulator pressure must be maintained at about 20,000 psi, even as the fuel is continually injected from the accumulator  12  into a plurality, e.g., four, six, or eight, engine cylinders. 
     High pressure fuel is delivered via line  14  through check valve  16 , from the pumping chamber  18 . The chamber  18  is formed at least in part by a bore  20  in which a pumping plunger  22  can reciprocate, in a manner well known in this field of technology. The plunger is directly driven by a rotating cam  24  having a cam profile  26 . While the plunger retracts, thereby enlarging the available volume in pumping chamber  18 , fuel is supplied via inlet passage  28  through check valve  30 . When the plunger  22  advances, the fuel in chamber  18  is delivered to the accumulator  12 . 
     Fuel is supplied through the inlet passage  28  at a pressure preferably in the range of about 200-300 psi, into high pressure pumping chamber  18 , by a positive displacement transfer pump  35 , preferably including a transfer pumping chamber  32  and associated pumping plunger or piston  34 . In practice, and as described more fully below, the system  10  would have a high pressure pumping chamber  18  formed by a plurality of pumping plungers and their respective bores, but only a single transfer pumping chamber  32  adapted to supply all the high pressure pumping bores. The transfer pump piston  34  is driven directly by a rotating cam  38  having a cam profile  36  which is different from, but has a pre-established timing relationship with, the cam profile  26 . 
     The fuel supplied to the transfer pump chamber  32  via passage  42  through check valve  44 , can be pre-metered, such as by a solenoid driven valve  40 . Electromagnet  46  is energized or de-energized via leads  48 , to retract or advance valve member  50 , away from or against valve seat  52 . This admits or blocks the flow of fuel from the low pressure supply line  56  from the fuel tank supply pump (not shown), through passage  54  into passage  42 . This fuel is typically at a pressure of less than about 20 psi, preferably 10-15 psi. The fuel supply at this low pressure can be considered a reservoir. 
     The valve  40  is relatively slow in operation, but as a consequence, is relatively accurate in the quantity of fuel that can be metered into chamber  32 . The quantity of metered fuel from valve  40  can be regulated according to the demand on the engine, e.g., during acceleration, in a well known manner. (For example by an ECU such as described in U.S. Pat. No. 5,103,792, “Processor Based Fuel Injection Control System”, the disclosure of which is hereby incorporated by reference.) The volume of the chamber  32  and the pre-load on the associated piston return spring  58 , assure that any potentially desirable quantity of metered fuel can be received in chamber  32  for delivery through passage  28 . 
     Regardless of the specific manner of supplying feed fuel to the pumping chamber, the components and flow paths which perform this function can be collectively referred to as a feed supply train. 
     The cam  24  and cam  38  are in rigid, fixed relation to each other, forming an actuating lever  60  which automatically coordinates the phasing of the relationship between the plunger  22  and the piston  34 . Such phasing can be understood with reference to FIGS. 2 and 3. The upper portions of FIGS. 2 and 3 represent the transfer pump chamber  32  and piston  34  as controlled by cam profile  36 , whereas the lower portions represents the high pressure pumping chamber  18  and plunger  22  as controlled by cam profile  26 . 
     Point A on the profile  26  corresponds to the cam nose, or peak displacement of plunger  22 , at zero rotation angle of cam  24 , and point B on the profile  36  corresponds to the minimum displacement of piston  34  at zero rotation angle of cam  38 . The complete cycle of one profile  26  from A to Al and one profile  36  from B to B 1 , is represented along a scale of zero to 100 per cent. As the piston  34  follows the upslope of portion  62  of profile  36 , the fuel in chamber  32  is discharged into the chamber  18 , because the plunger  22  is retracting as it follows the downslope of cam portion  64 . The charge of fuel delivered to chamber  18  is thus commensurate with (and preferably equal to) the pre-metered quantity to chamber  32 . During steady state operation, the quantity of fuel in chamber  32  delivered to chamber  18 , only partially fills chamber  18 , as shown at 40 per cent scale. Chamber  18  does not fully expand, but rather reaches an intermediate limit at about 20 per cent scale (point  76 ), and remains at that limit until just past 60 per cent scale. The downslope, minimum, and upslope portions defined by segments  76  to  66 ;  66  to  68 ; and  68  to  78  do not influence the fuel volume ultimately charged in the chamber  18 . At point  78 , the plunger  22  advances through chamber  18  to develop the high pressure for delivery to the accumulator  12 . 
     As the piston  34  follows the downslope portion  61  of cam profile  36 , the chamber  32  expands to receive the metered supply of fuel via valve  40 . This quantity is delivered during a relatively long period of time during which the high pressure plunger  22  is delivering fuel from the chamber  18  to accumulator  12 , as a result of the upslope on portion  74  of profile  26 . The quantity of fuel supplied to chamber  32  is calculated by an on-board computer or regulator (not shown) depending on the desired fuel delivery to the engine and the desired accumulator pressure. The maximum displacement of the transfer pump piston  34  is slightly smaller than the maximum displacement of the high pressure plunger  22  (e.g., 10 per cent less) in order to avoid hydraulic lock and to protect the pump components from mechanical overstress. 
     It can be appreciated from FIGS. 1 and 2, that the normal pumping rate of cam profile  26  is relatively low (i.e., a gradual and relatively long upslope  74  along almost 40 per cent of scale), which minimizes both hydraulic and acoustic noise. This has been achieved because the necessary quantity of fuel is delivered to the pumping chamber  18 , over a relatively short time period (i.e., during less than about 20 per cent of scale on the steep downslope  64 ). 
     In the preferred implementation, an inexpensive, easily controlled transfer pump arrangement is capable of delivering a metered quantity at high pressure over a short time interval. The quantity is controllable by the use of an inexpensive valve  40 , because the time available for metering quantity, is relatively long, i.e., the full length of profile  61 . Yet the metered quantity can be delivered to charge chamber  18  at a pressure of, for example, 200-300 psi, thus requiring only a short delivery time interval. This is in contrast to conventional transfer pumps, which typically operate at less than about 15 psi, and thus require a considerably longer time interval to charge the same quantity. 
     The capability to charge the high pressure pumping chamber  18  during a short interval (e.g., within about 10-20 per cent of scale rotation of cam  26  for steady state operation), not only permits the use of a long, gradually ascending profile  74  for the driving of the high pressure plunger  22 , but further permits accommodation of a dual ascending rate. This is shown in FIG. 2, as a short, rapidly rising profile  70 , between points  68  and  72 , followed by the longer, lower rate portion  74 . The slope of profile portion  70  is preferably at least twice as steep as that of profile portion  74 . During steady state, the chamber  18  does not completely fill, so the high rate portion  70  is not utilized. This is represented by the dashed line extending between points  76  and  78 , whereby the plunger  22  “floats” for a duration of about 40 per cent scale. 
     During demand for faster accumulator pressure increase associated with acceleration, the valve  40  admits a higher quantity of fuel to chamber  32 , which corresponds to a longer duration on portion  64  of profile  26 , almost to point  66 , thereby nearly filling chamber  18 . This situation is explained with respect to FIG.  3 . The plunger  22  then floats along a “flat” transition profile between  66  and  68 , before quickly rising along portion  70  and then continuing the pumping action along the “steady state” slope  74  between points  72  and A 1 . The portion  70  used for acceleration, preferably spans a duration of about 10-15 per cent of scale. Even in the transient operation depicted in FIG. 3, the plunger  22  floats for a duration of about 20 per cent scale, on an oil film indicated at the arrows. The plunger  22  thus releases the cam force loading the shoe  116   b  for a certain time period to allow the roller  116   b  to replenish the oil film inside of the shoe. Preferably, the volumetric charging rate into chamber  22  is at least 50% faster than the acceleration pumping rate due to cam profile portion  70 . 
     The point  78  at which the roller begins the pumping phase on profile  26  by actuating plunger  22  inwardly, depends on the volume of fuel transferred from chamber  32  into the pumping chamber  18 . This volume is commensurate with, and preferably substantially equal to, the volume of fuel metered by valve  40  during the intake stroke of piston  34  along transfer cam portion  61 . The charge of fuel delivered by the piston stroke along cam portion  62  through path  28 , is predictably allocated in the pumping chamber so each of the plunger bores receives approximately the same amount of fuel during charging. 
     FIGS. 4 and 5 show longitudinal and cross sectional views of a preferred embodiment  100  for implementing the inventive features described above. A pump shaft housing  102  has a central cavity  104  in which a drive shaft  106  is supported for rotation. A stationary and rotationally fixed body  108  is situated in part within housing  102  and coaxially aligned with the shaft. Fixed head  110  is secured to housing  102  and has a low pressure fluid handling portion  112  including low pressure supply and leak off channels handling portion  112  including low pressure supply and leak off channels  112   a ,  112   b , and a roller shoe support hub portion  114 . The hub  114  lies within housing  102  and surrounds body  108 . The shaft housing  102  and head  110  together can be considered as defining the housing of the pump assembly. 
     The associated roller assembly  116  is situated concentrically inside the cam ring  118 , which rotates as a result of fixation to the shaft  106 . Four radially extending, orthogonally oriented bores  20  in the body  108  contain a respective four reciprocable plungers  22  which cause the respective pumping volumes to expand and contract. The pumping chamber  18  is formed at the intersection of the bores  20  in central cavity  120  of body  108 . It should be understood that the pumping chamber as shown is a volume defined cooperatively by portions in the central cavity and inner ends of the bars, that is pressurized by the simultaneous actuation of the plungers. A high pressure outlet fitting  122  is fixedly supported within the cavity  120 , and an axially slidable control valve  124  is supported within the fitting  122 . The cam ring  118  surrounds the pumping plungers  22  and, in a manner well known in this field, a cam pumping profile  26  along the inner circumference of the ring, cooperates with cam rollers  116   a , and associated shoes  116   b , to reciprocate the plungers  22 . This overall arrangement is analogous to that described in U.S. Pat. Nos. 5,215,449 issued Jun. 1, 1993, U.S. Pat. No. 5,688,110 issued Nov. 18, 1997 (the disclosure of which is hereby incorporated by reference). 
     The outer circumference of the cam ring  118  also provides a cam profile  36 , for maintaining rolling contact with an outer roller  126  which causes reciprocation of the piston  34  in the transfer pump  35 . The cam ring as shown in FIG. 5, is rotatable counterclockwise and is depicted at the “zero” angle of rotation on the scale for profile  26 , wherein the high pressure pumping roller  116   a  is on the nose of the pumping profile, corresponding to point A in FIG.  2 . The transfer roller  126  is at the lowest point B of the cam profile  36  as shown in FIG.  2 . Rotation of the cam ring until point A 1  arrives at the former location of point A, corresponds to one pumping cycle. The total travel ways, for example, as “100 per cent of a pumping cycle”, or as the angular displacement of the cam ring  118  which in the illustrated embodiment is 45 degrees. Clearly, for a different number of plungers or actuation frequency, 100 per cent of a pumping cycle could correspond to a different angular displacement, such as 60 degrees or 90 degrees. 
     In the embodiment of FIG. 4, the roller  126  for the transfer pump is in vertical alignment with the rollers  116   a  for the high pressure pump, but this is not necessary. Furthermore, only one transfer pump  35  serves all high pressure pumping bores  20 . Fuel is transferred to all pumping bores simultaneously during a relatively short portion of the pumping cycle, and thereafter all plungers  22  are driven inwardly simultaneously during a longer portion of the pumping cycle. 
     In general, advantageous use of the pump as described above with respect to FIGS. 1-5, can be realized within the range of parameters shown in Table 1: 
     
       
         
               
               
               
               
             
               
               
               
               
               
             
           
               
                   
                 TABLE 1 
               
               
                   
                   
               
               
                   
                                 Feature 
                 Scale Duration 
                    Numeric ID 
               
               
                   
                   
               
             
             
               
                   
               
             
          
           
               
                   
                                  Transfer Cam Profile 
                 100 
                 percent 
                 36 
               
               
                   
                 intake portion 
                 &gt;50 
                 percent 
                 61 
               
               
                   
                 discharge portion 
                 &lt;50 
                 percent 
                 62 
               
               
                   
                 Pumping Cam Profile 
                 100 
                 percent 
                 26 
               
               
                   
                 nose portion 
                 &lt;5 
                 percent 
                 A 
               
               
                   
                 charging portion 
                 20-30 
                 percent 
                 64 
               
               
                   
                 flat portion 
                 10-30 
                 percent 
                 66 to 68 
               
               
                   
                 acceleration portion 
                 10-20 
                 percent 
                 70 
               
               
                   
                 steady state portion 
                 30-60 
                 percent 
                 74 
               
               
                   
                   
               
             
          
         
       
     
     The preferred implementation of the cooperation between the transfer pump  35  and the control valve  124  for charging and discharging the high pressure pumping chamber  18 , is shown schematically in FIG.  6 . The transfer pump  35  is particularly well suited for rapidly transferring a metered volume of fuel to the inlet check valve  132 , to charge the pumping chamber during only a short duration of the retraction of the pumping plungers  22  along cam profile portion  64  (e.g., &lt;20 per cent of scale during steady state maximum fuel delivery, as shown in FIG.  2 ). Also, the transfer pump phasing offset and reduction in the number of seals between the transfer pump and the high pressure chamber permits the transfer pump itself to initially charge the accumulator to a pressure of about 200-300 psi, before the high pressure pumping takes over. This can substantially reduce the cranking time and reduce the pressurized response time whenever required for a cold engine. As shown with greater particularity in FIG. 6, the transfer pump roller  126  actuates transfer piston  34 , whereby fuel is delivered via low pressure supply line  112   a , into the exterior groove in pump body  108 , whereupon the fuel is delivered via inlet passage  130  to check valve  132 . Check valve  132  opens during the intake phase of pumping operation, thereby delivering fuel into the pumping chamber  18  whereupon, as the plungers (not shown in FIG. 6) are actuated radially inwardly, the inlet check valve  132  closes. The control valve  124  is normally spring biased to prevent passage of fuel through high-pressure passages  134  during the intake phase, but during the pumping phase, the valve  124  opens, so that high pressure fuel is delivered via high pressure passages  134  and valve cavity  136 , to the discharge port  138  (see FIG.  4 ). 
     FIGS. 7 and 8 show another embodiment of a high pressure supply pump  200  with simultaneously directly actuated plungers, which shares many features of the pump shown in FIGS. 4-6, except that the feed technique can be different. The actuating cam  202  only has an internal cam profile, without the external profile for actuating a transfer pump. Feed fuel is delivered via modulated pressure inlet  204  through paths  206  in housing or housing cover and  207  in central body  208 , to the inlet chamber  210 , located on the central axis on one side of the pumping plunger plane, from which feed fuel can flow into the pumping chamber  212  upon retraction of the inlet check valve  214 . During the pumping phase of operation, the fuel at high pressure leaves the pumping chamber  212  through the respective discharge passages  216  as the discharge or outlet check valve  218  opens and the inlet check valve closes. The fuel flows over the control valve  220  and is discharged from the pump housing via the outlet  222 . 
     Leak-off grooves  224  are provided in the pumping plunger bores  20 , for return of fuel via  230  to the fuel tank at low pressure. The outer surface  226  of a portion of the outlet fitting  228  is spaced from the body  208 , to provide a portion of the leak-off flow path. 
     The simultaneous pumping permits more pumping cycles per drive shaft revolution, than is available in a sequentially actuated pump, and therefore a greater quantity of fuel can be provided at high pressure through the outlet, during a given time period. 
     It can be appreciated that the engine lube oil introduced through orifice  232  and delivered through path  234  in the housing  236  is utilized for lubricating the sliding surfaces of the pump, i.e., the interaction of the internal cam  202  with the shoe and roller  238  and plunger outer surfaces  240 .