Abstract:
A diaphragm pump which overcomes the problem of diaphragm failure due to overfill of the oil transfer chamber and the inability to self-prime. A notch is provided in the upper portion of the surface of the cylinder so that air can be forced back to the reservoir. In addition, the bias spring connected to the diaphragm and supported by the piston is made stiff with a spring constant that produces a bias pressure that can overcome abnormal suction pressures.

Description:
FIELD OF THE INVENTION  
         [0001]    The present invention relates generally to an improved diaphragm pump, and more specifically, to an improved diaphragm pump for use under a condition where the hydraulic fluid side of the diaphragm is primed and the pumping side of the diaphragm is in a relatively high vacuum state and another condition where the hydraulic fluid side of the diaphragm is not primed.  
         DESCRIPTION OF THE ART  
         [0002]    The known rotary-operated, oil-backed/driven diaphragm pump is a high-pressure pump inherently capable of pumping many difficult fluids because in the process fluid, it has no sliding pistons or seals to abrade. The diaphragm isolates the pump completely from the surrounding environment (the process fluid), thereby protecting the pump from contamination.  
           [0003]    In general, a diaphragm pump  20  is shown in FIG. 1. Pump  20  has a drive shaft  22  rigidly held in the pump housing  24  by a large tapered roller bearing  26  at the rear of the shaft and a small bearing (not shown) at the front of the shaft. Sandwiched between another pair of large bearings (not shown) is a fixed-angle cam or wobble plate  28 . As the drive shaft turns, the wobble plate moves, oscillating forward and back converting axial motion into linear motion. The three piston assemblies  30  (only one piston assembly is shown) are alternately displaced by the wobble plate  28 . As shown later, each piston is in an enclosure including a cylinder such that the enclosure is filled with oil. A ball check valve  32  in the bottom of the piston/cylinder assembly  30  functions to allow oil from a reservoir  27  (wobble plate  28  is in the reservoir) to fill the enclosure on the suction stroke. During the output or pumping stroke, the held oil in the enclosure pressurizes the back side of diaphragm  34 , and as the wobble plate moves, causes the diaphragm to flex forward to provide the pumping action. Ideally, the pump hydraulically balances the pressure across the diaphragm over the complete design pressure range. As discussed later, in actual practice this is not the case for all situations for known pumps. In any case, each diaphragm has its own pumping chamber which contains an inlet and an outlet check valve assembly  36 ,  37  (see also FIG. 2). As the diaphragm retracts, process fluid enters the pump through a common inlet and passes through one of the inlet check valves. On the output or pumping stroke, the diaphragm forces the process fluid out the discharge check valve and through the manifold common outlet. The diaphragms, equally spaced 120° from one another, operate sequentially to provide constant, virtually pulse-free flow of process fluid.  
           [0004]    In more detail, a portion of a diaphragm pump  20  is shown in cross-section in FIG. 2. The diaphragm  34  is held between two portions  38 ,  40  of housing  24 . Diaphragm  34  separates the pump side from the oil-filled, hydraulic drive side of the pump. On the drive side, a drive piston assembly  30  including a diaphragm plunger  42  are contained within the oil filled enclosure which functions as a transfer chamber  44 . A plurality of check valves  32  in piston  46  separate transfer chamber  44  from the oil reservoir (not shown). Wobble plate  28  (not shown in FIG. 2) contacts pad  48  to drive piston  46 . Arrow  49  indicates the general direction of movement of the cam or wobble plate. When the piston and diaphragm have finished the forward or pumping stroke, the end  50  of piston  46  is at top dead center (TDC). When the piston and diaphragm have retracted in the suction stroke, the end  50  of piston  46  is at bottom dead center (BDC).  
           [0005]    Piston  46  reciprocates in cylinder  47 . Piston  46  has a sleeve section  52  which forms the outer wall of the piston. Sleeve section  52  includes a sleeve  54  and an end portion  56  at the end having pad  48  which is contact with the wobble plate. Within sleeve  54  is contained a base section  58 . Base section  58  includes a first base  60  which is in contact with end portion  56  and includes seal elements  62  for sealing between first base  60  and sleeve  54 . Base section  58  also includes second base  64  at the end opposite of first base  60 . Connecting wall  66  connects first and second bases  60  and  64 . Piston return spring  68  is a coil spring which extends between first base  60  and diaphragm stop  70  which is a part of the pump housing  24 . Valve housing  72  is contained within base section  58  and extends between second base  64  and end portion  56 . Seals  74  provide a seal mechanism between valve housing  72  and connecting wall  66  near second base  64 .  
           [0006]    The end  76  opposite end portion  56  of sleeve portion  52  is open. Likewise, the end  78  of valve housing  72  is open. Second base  64  has an opening  80  for receiving the stem  82  of plunger  42 .  
           [0007]    Diaphragm plunger  42  has the valve spool  84  fitted within valve housing  72  with the stem  82  extending from the valve spool  84  through opening  80  to head  86  on the transfer chamber side of diaphragm  34 . Base plate  88  is on the pumping chamber side of diaphragm  34  and clamps the diaphragm to head  86  using a screw  90  which threads into the hollow portion  92  of plunger  42 . Hollow portion  92  extends axially from one end of plunger  42  to the other end. Screw  90  is threaded into the diaphragm end. The piston end of hollow portion  92  is open. A plurality of radially directed openings  94  are provided in stem  82 . A bias spring  96  is a coil spring and extends between second base  64  and valve spool  84 . A valve port  98  is provided in the wall of valve housing  72 . A groove  100  extends in connecting wall  66  from the furthest travel of valve port  100  to end portion  56 . A check valve  102  is formed in end portion  56  in a passage  104  which is fluid communication with the reservoir (not shown). Thus, there is fluid communication from the reservoir (not shown) through passage  104  and check valve  102  via groove  100  to valve port  98 . When the valve is open, there is further communication through the space in which coil spring  96  is located and then through one of the plurality of radial openings  94  and through the axial hollow portion  92  of plunger  84 . There is further fluid communication from the hollow portion  92  through the other radially directed openings  94  to various portions of transfer chamber  44 . The hollow passage  92 , along with the radially directed openings  94  provide fluid communication from the portion of transfer chamber  44  near diaphragm  34  to the portion of transfer chamber  44  within the valve housing  72  of piston  30 . The transfer chamber also includes the space occupied by piston return spring  68 .  
           [0008]    On the pump side of diaphragm  34 , there is an inlet check valve assembly  36  which opens during the suction stroke when a vacuum is created in pumping chamber  106 . There is also a check valve  37  which opens during the pumping or output stroke when pressure is created in pumping chamber  106 .  
           [0009]    FIGS.  3 ( a )-( f ) illustrate operation of the conventional pump  20  under normal, standard operating conditions using a conventional bias spring  96 . Typical pressures are shown. Typical vector directions for the cam or wobble plate (not shown in FIGS.  3 ( a )-( f )) are shown. Suction is less than 14.7 psia. Output pressure is greater than 14.7 psia. The pressure differential across diaphragm  34  is set at about 3 psi.  
           [0010]    With reference to FIG. 3( a ), the suction stroke begins at the end of the pumping stroke. For the conditions assumed, pressure in the pumping chamber immediately drops from what it was at high pressure, for example, 120 psia to 10 psia. Pressure in the hydraulic transfer chamber is 13 psia which is less than the 14.7 psia in the reservoir. The piston  30  is at top dead center and begins moving toward bottom dead center. Bias spring  96  momentarily moves plunger  42 , and particularly valve spool  84 , to the right to open port  98 . Because pressure in the transfer chamber is less than the pressure in the reservoir, check valve  32  opens and oil flows from the reservoir to the transfer chamber to appropriately fill it with oil which had been lost during the pumping stroke previous. That is, under the pressure of the pumping stroke oil flows through somewhat loose tolerances of the parts of the piston so that some of the oil flows from the transfer chamber back to the reservoir. Thus oil needs to be refilled in the transfer chamber during the suction stroke so that there is enough oil to efficiently provide pressure during the next pumping stroke.  
           [0011]    [0011]FIG. 3( b ) shows the configuration at mid-stroke. The slight suction in the pumping chamber (shown to be 10 psia), holds diaphragm  34  and spool  84  to the left while piston  30  moves to the right, thereby shutting off port  98 . Since pressures are nearly equal and diaphragm  34  moves right with piston  30 , the pumping chamber fills with process fluid.  
           [0012]    As shown in FIG. 3( c ), process fluid continues to fill as diaphragm  34  moves right. Valve port  98  remains shut. Very little leakage of oil occurs from the reservoir (not shown) to transfer chamber  44 , since pressures are nearly equal. Thus, both sides of the diaphragm fill properly.  
           [0013]    When piston  30  reaches bottom dead center, the suction stroke is completed and the output or pumping stroke begins as shown in FIG. 3( d ). Pressure in the transfer chamber immediately increases, for example, from 13 psia to 123 psia. Likewise, pressure in the pumping chamber immediately increases, for example, from 10 psia to 120 psia. The wobble plate begins moving piston  30  to the left which causes the build-up of pressure. Check valves  32  close. Diaphragm  34  moves in volume tandem with the oil and process fluid left with the piston to push (pump) process fluid out.  
           [0014]    At mid-stroke as shown in FIG. 3( e ), there is continued output. Some oil leakage past the tolerances between piston and cylinder may move valve spool  84  of diaphragm plunger  42  to the right to open valve port  98 . Check valves  32 , however, are closed, thereby locking the oil in transfer chamber  44 , except for leakage.  
           [0015]    The output stroke finishes with the configuration shown in FIG. 3( f ). The filled transfer chamber  44  pushes diaphragm  32  to the left dispensing process fluid as it moves. Normal operation as shown in FIGS.  3 ( a )-( f ) causes little stress on diaphragm  32 .  
           [0016]    A problem with conventional diaphragm pumps, however, is an unexpected diaphragm rupture under certain operating conditions. The, diaphragm can fail much sooner than normal, or more frequently, may fail sooner than other pump components. A failure contaminates the process lines with drive oil. The operating condition which most often causes failure is a high vacuum inlet with a corresponding low outlet pressure. This is an expected occurrence in a typical pumping system when the inlet filter begins to plug. In that case, the plugging requires high vacuum to now pull process fluid through the filter. At the same time, the lowering of process fluid volume pumped drops the outlet pressure. This creates a situation where a high suction on the pumping side lowers the pressure during the suction stroke on the transfer chamber side so that the transfer chamber essentially “asks for more fill fluid” and, consequently, in-flowing oil overfills the transfer chamber and does so without a corresponding high pressure to push oil out during the pumping or output stroke to counter-balance. The overfill of oil “balloons” the diaphragm into the fluid valve port until the diaphragm tears. Additionally, with a high-speed, reversing, vacuum/pressure pump such as this apparatus, the high-speed valve closings create tremendous pressure spikes, called Jaukowski shocks. The spikes can consist of fluid pressure or acoustical waves and harmonics of both. These pressure spikes can “call for” oil fluid flow into the drive piston when that should not be happening. Again, this can cause overfill and lead to the diaphragm failure. FIGS.  4 ( a )- 4 ( f ) are provided to illustrate the overfill failure mode.  
           [0017]    In FIG. 4( a ), the suction stroke begins. Since it is assumed that the inlet side for the process fluid is plugged or blocked off, only a low pressure was created during the output stroke. That is, the pressure in the pumping chamber  106  was, for example, 14 psia and goes to 10 psia as it did in FIG. 3( a ). The suction, however, quickly increases the vacuum so that pressure in the pumping chamber  106  drops further to, for example, 3 psia as shown in FIG. 4( b ). The diaphragm  34  and plunger  42  stay too far left keeping valve port  98  closed and bias spring  96  somewhat compressed. There is only momentary oil flow through check valves  32 , valve port  98  and the various passageways in stem  82 .  
           [0018]    At mid-stroke of the suction stroke as shown in FIG. 4( b ), any diaphragm movement right causes a higher vacuum in pumping chamber  106  which tends to hold diaphragm  34  and plunger  42  to the left, while piston  46  moves to the,right. Valve port  98  is shut off, but nevertheless because of the lower pressure, for example, 6 psia, being developed in transfer chamber  44 , there is oil leakage due to the tolerances in the system from the reservoir (not shown) to transfer chamber  44 . The weak bias spring  96  in the conventional diaphragm pump allows plunger  42 , and particularly valve spool  84 , to stay too far left and allow the lower pressure in transfer chamber  44  to develop and continue.  
           [0019]    As shown in FIG. 4( c ), at the end of the intake or suction stroke, the plunger  42  and diaphragm  34  remain too far left, and the low pressure in transfer chamber  44  continues to cause leakage and after many strokes like this, transfer chamber  44  gets overfilled with oil prior to starting the output stroke.  
           [0020]    The configuration at the beginning of the output stroke is shown in FIG. 4( d ). Piston  46  starts to move left. Since there is low pressure in the pumping chamber  106 , pressure does not build in transfer chamber  44  until later in the output stroke.  
           [0021]    As shown at mid-stroke in FIG. 4( e ), the overfilled oil transfer chamber  44  moves diaphragm  34  and valve spool  84  to the left at the same rate. When base plate  88  and diaphragm  34  approach wall  108  on the pumping side of the pump, pressure finally rises in transfer chamber  44 . The short time in which there is pressure greater than 14.7 psia, which is the pressure in the reservoir, is not enough time to allow oil leakage back from transfer chamber  44  to the reservoir to balance flow leakage during the suction stroke. Hence, the diaphragm  34  distorts due to the oil overfilling in transfer chamber  44 . The weak spring  96  is compressed.  
           [0022]    The end of the output stroke is shown in FIG. 4( f ). Overfilled transfer chamber  44  pushes base plate  88  fully against wall  108  and diaphragm  34  stretches into the port of outlet check valve assembly  37 . A rapid rise in pressure in transfer chamber  44  at this time eventually causes diaphragm  34  to either cut on various surfaces it encounters or to burst. At this point, the pump fails. As a result, there can be contamination of process fluid remnants into piston assembly  30  and contamination of oil into the process fluid line.  
           [0023]    Thus, when a high vacuum (that is, a plugged filter or inlet valve shut off) exists on the pumping chamber side of the diaphragm, the diaphragm does not want to move with the piston. This would not ordinarily cause a problem, as the valve spool  84  and valve port  98  close. If this condition exists, however, for a long period of time, the leakage between the valve spool and the valve port plus the leakage between the piston and the housing combine to allow oil overfill in the transfer chamber. On the output stroke, the pressure must be high enough to re-expel leakage volume. It can expel, however, only around the piston and housing since the ball check valves  32  prevent any exiting through the valve port. Since the pump inlet is blocked and unable to pump much process fluid volume, pressure during process fluid outlet is low and/or only for part of the stroke. Empirically, it has been found that the outlet pressure must be more than 100 psig in order to “leak as much out as in”. If the pump does not leak as much out of the transfer chamber as it leaks in, then the added volume is powered by the drive piston until the diaphragm balloons and enters ports or crevices and causes rupture.  
           [0024]    Conventional pump  20  also has the problem that valve spool  84  can stick to burrs in particular at the edge of openings for valve ports  98 . In this type of situation, diaphragm  34  tends to wrap around base plate  88  thereby stressing and/or pinching the diaphragm material.  
           [0025]    Conventional pump  20  has the further problem of volumetric inefficiency. This occurs because there is not a large enough bypass leakage of oil (and air) around the piston to purge the air from the transfer chamber. Under this condition, efficiency decreases as more and more air accumulates within the transfer chamber. This decreased volumetric efficiency occurs because the piston repeatedly compresses and decompresses the excess of air caught in the transfer chamber. This causes more and more severe fluid pressure pulsation because air compressing changes the diaphragm stroke from pure sinusoidal form to almost a square form. A direct result of this is increased pressure fluctuation at the pump outlet,, an undesirable characteristic of a diaphragm pump.  
         SUMMARY OF THE INVENTION  
         [0026]    The present invention is directed to a diaphragm pump which receives drive power from a motor. The pump has a casing which houses a pumping chamber adapted to contain fluid to be pumped (process fluid), a transfer chamber adapted to contain hydraulic fluid (oil), and a hydraulic fluid reservoir. The pump has a diaphragm having a transfer chamber side and a pumping chamber side. The diaphragm is supported by the casing and is disposed between the pumping chamber and the transfer chamber and is adapted for reciprocation toward and away from the pumping chamber. The pump has a piston in a cylinder in the casing adapted for reciprocation of the diaphragm between a power stroke and a suction stroke.  
           [0027]    The cylinder forms a portion of the transfer chamber. The piston moves longitudinally in the cylinder with the cylinder when the pump is oriented so that the cylinder is generally horizontal having a surface with an upper portion. A wobble plate and a first spring cooperate to reciprocate the piston. The wobble plate is driven by the motor. The first spring is compressible between the housing and the piston. A second spring urges the diaphragm away from the pumping chamber with a first end of the second spring connected with the diaphragm and a second end of the second spring supported by the piston for movement therewith. A fluid communication path for the hydraulic fluid is formed between the hydraulic fluid reservoir and the transfer chamber. A valve in the fluid communication path allows selectively flow of hydraulic fluid from the hydraulic fluid reservoir to the transfer chamber when the valve is open. A vent is formed in the upper portion of the surface of the cylinder. In this way, air in the transfer chamber is forced from the transfer chamber throughout the vent in the cylinder so as so enhance the quality of the fluid remaining in the transfer chamber and to self-prime the pump.  
           [0028]    In this way, the present invention discloses a novel diaphragm pump that “spits” out small amounts of trapped air and oil through the vent on each cycle of the pump. It does this only at a point in the stroke where no large shock pressures are occurring. Having only non-compressing oil in the cylinder provides “solid” displacement to enhance metering of oil, volumetric efficiency, and outlet pressure stability of the pump. Removing air prevents the problems caused by accumulated air entrapment, including the inability to self-prime. This simplifies final assembly, final test, and user operation. The present invention maintains the biased oil drive as described in U.S. Pat. No. 3,775,030. The present invention, however, discloses use of a stiff bias spring. In this way, at high vacuum conditions, the bias spring keeps drive oil pressure above its vapor pressure, which prevents oil cavitation, and (2) the bias spring overcomes suction forces in the pumping chamber and prevents oil overfill in the transfer chamber (so the diaphragm does not fail).  
           [0029]    Thus, the improvements disclosed herein optimize durability and efficiency for a diaphragm pump. 
       
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0030]    [0030]FIG. 1 is a perspective view of a conventional diaphragm pump;  
         [0031]    [0031]FIG. 2 is a partial cross-sectional view of a conventional diaphragm pump;  
         [0032]    FIGS.  3 ( a )- 3 ( f ) are partial cross-sectional views of a conventional diaphragm pump illustrating normal conditions;  
         [0033]    FIGS.  4 ( a )- 4 ( f ) are partial cross-sectional views of a conventional diaphragm pump illustrating a high vacuum condition resulting in diaphragm failure;  
         [0034]    [0034]FIG. 5 is a partial cross-sectional view of a diaphragm pump in accordance with the present invention;  
         [0035]    [0035]FIG. 6 is a partial cross-sectional view of a first alternate embodiment;  
         [0036]    [0036]FIG. 7 is a partial cross-sectional view of a second alternate embodiment;  
         [0037]    [0037]FIG. 8 is an exploded, cross-sectional view of a piston/cylinder assembly;  
         [0038]    FIGS.  9 ( a )- 9 ( f ) are partial cross-sectional views of a diaphragm pump illustrating operation with a high spring constant bias spring;  
         [0039]    [0039]FIG. 10 is a graph illustrating a weak conventional bias spring and a strong bias spring in accordance with the present invention;  
         [0040]    [0040]FIG. 11 is a graph which illustrates a range of spring constants for bias springs in accordance with the present invention; and  
         [0041]    FIGS.  12 ( a )- 12 ( f ) are partial cross-sectional views of a diaphragm pump having an air-expelling notch and illustrating self-priming. 
     
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0042]    The present invention is an improvement to the conventional diaphragm pump described above. Like parts are designated by like numerals. Improved parts are distinguished and described. It is understood that the improved parts lead to a synergistic improvement of pump performance and durability.  
         [0043]    With reference to FIG. 5, the present invention is embodied in pump  110 . Housing  112  comprises portions  38 ,  114  which are similar to portions  38 ,  40  of housing  24 . Portion  114  includes a vent with a form of a notch  116  formed in the upper portion  118  of the surface of cylinder  120 , which is similar to cylinder  47 . Notch  116  provides fluid communication between transfer chamber  44  and the oil reservoir (not shown). Although notch  116  is shown to extend from beyond the right end of piston  46  in cylinder  120  when piston  46  is as far right as it can travel, namely, when base plate  88  contacts wall  122  of housing portion  38 , the preferred embodiment has the notch extending just past the halfway forward travel of the piston. Thus the piston will “valve off” the notch passage during the final half of the output stroke and the first half of the suction stroke. The notch will open to expel air and oil just before midpoint of the suction stroke and stay open till just past midpoint of the output stroke. This has empirically proven to provide the required easy priming while minimizing leakage. Notch  116  extends to the left to the end  124  of housing portion  114  where it opens to the oil reservoir.  
         [0044]    It is further noted that pump  110  has a significantly stiffer bias spring  126 . The combination of the significantly stiffer bias spring  126  and notch  116  leads to virtual elimination of diaphragm failure when a high vacuum condition develops on the pumping side of the diaphragm and also leads to reduction of air in the hydraulic fluid in transfer chamber  44  and, consequently, allows pump  110  to achieve self-priming.  
         [0045]    A first embodiment of the present invention is shown in FIG. 6. Pump  127  shows a notch  128 , similar to notch  116 , except notch  128  does not extend all the way to end  124 . Rather, a radially extending passage  130  in said housing portion  114  extends from the end of notch  128  near end  124  to an O-ring groove  132 . O-ring  134  is provided in groove  132 .  
         [0046]    O-ring  134  in groove  132  functions as a check valve. Whenever sufficient pressure exists in transfer chamber  44 , the pressure will slightly open O-ring  134  from passage  130  to allow air/oil to be expelled into the reservoir (not shown). With this embodiment, fluid flows only out through notch  128 , passage  130  and the check valve of O-ring  134  and groove  132 , as opposed to two-way flow through notch  116  of pump  110 .  
         [0047]    A second alternative embodiment of the present embodiment is shown in FIG. 7. Pump  129  shows a passage  131  extending from the upper portion  118  of cylinder  120 . Passage  131  extends through wall  133  of portion  135  of housing  137 . Passage  131  provides fluid communication between transfer chamber  44  and the hydraulic fluid reservoir. Preferably, passage  131  extends radially and vertically. Preferably also, passage  131  is located just past the halfway forward travel of piston  46 . Thus, piston  46  will “valve off” the passage during the final half of the output stroke and the first half of the suction stroke. The passage will open to expel air and oil just before the midpoint of the suction stroke and stay open until just past the midpoint of the output stroke. Thus, passage  131  provides similar function as notch  116 .  
         [0048]    Another feature of the present invention which is relevant to all embodiments is shown in FIG. 8. Valve housing  136  includes a circumferential groove  138  which is axially located so as to intersect with valve port  140 . Without groove  138 , there is a chance of a burr being formed when the radial valve port opening is manufactured. If there is a burr present, then valve spool  84  can get caught on the burr so that the spool sticks. In this case, the diaphragm  34  may wrap around base plate  88  and become stressed and/or pinched. By forming the circumferential groove  138 , the possibility of such a burr is eliminated.  
         [0049]    In operation, a design configuration wherein a pump in accordance with the present invention has a stiff bias spring  126 , as distinguished from a weak bias spring  96 , is described with respect to FIGS.  9 ( a )-( f ). A weak bias spring  96  of a conventional pump is distinguished from a stiff bias spring  126  in FIG. 10.  
         [0050]    [0050]FIG. 10 is a graph which shows spring length in inches along the X-axis. On the left side along the Y-axis, the graph is calibrated for force in pounds which the piston exerts on the diaphragm. Along the right side for the Y-axis, an effective pressure at the diaphragm in pounds per square inch (psi) is provided. In the conventional pump, it is known from U.S. Pat. No. 3,775,030, that a small over-pressure, for example,  3  psi, should be provided in the transfer chamber  44  in order for the pump to work properly under normal conditions. As consequence, the conventional thinking has been to provide a weak spring so that the over-pressure maintained by the bias spring does not differ too greatly from  3  psi for various spring lengths during the compression of normal operation. A spring constant for a typical spring is shown as line  140  in FIG. 10. However, as discussed above with respect to FIGS.  4 ( a )- 4 ( f ) the conventional pump has the problem of the diaphragm  34  failing if the line providing process fluid to the pump becomes plugged, such as when a filter gets dirty. Thus, with respect to the present invention, two reference points were considered. A first reference point occurs when valve port  121  in FIG. 5 or valve port  98  in FIG. 2 just turns off or is closed. At the point at which valve port  98  just turns off, the bias spring should counteract fluid suction on the fluid pumping side adequately to prevent the suction from holding the diaphragm to that side and thereby allowing unwanted oil to fill into the transfer chamber. The minimum, of course, is zero since clearly a negative pressure would constantly call for more oil in the transfer chamber and be undesirable. Experience with the conventional pump as discussed above has shown that  3  psi works well. Somewhat greater, up to 4 psi or so, is acceptable. Therefore, a range of zero-4 psi is appropriate. Reference point  1  is shown at numeral  142  in FIG. 10.  
         [0051]    The second reference point occurs when transfer chamber  44  has filled with oil to its maximum, that is, when base plate  88  contacts wall  108  as shown in FIG. 4( f ). The second reference point is shown at numeral  144 . For weak spring  140 , the pressure at valve shut off reference point  142  is slightly greater than 3 psi and at maximum overfill reference point  144  the pressure is about 4 psi. Conventionally, this has been the design for bias spring  96 . In order to solve the problem of diaphragm failing for a high vacuum condition in the pumping chamber of the pump, however, it was determined that it was necessary to) approximately satisfy reference point  1  with respect to normal operating conditions, and with respect to the condition of high vacuum, it was determined that the spring should provide a pressure in transfer chamber  44  of about 10.5 psi as shown at numeral  146  in FIG. 10, which does not allow a large pressure differential between the reservoir and the transfer chamber. The reservoir is atmospheric, or essentially 14.7 psi. These two reference points when connected by a straight line then determine the spring constant for the improved pump.  
         [0052]    FIGS.  9 ( a )- 9 ( f ) illustrate operation with respect to a stiff spring of the type represented by line  148  in FIG. 10.  
         [0053]    FIGS.  9 ( a )- 9 ( f ) assume the stiff bias spring and a vacuum condition, that is, a plugged process line. FIGS.  9 ( a )- 9 ( f ) are similar to FIGS.  4 ( a )- 4 ( f ), except the weak bias spring is replaced by the stiff bias spring.  
         [0054]    In FIG. 9( a ), the suction stroke begins. Since the inlet for the process fluid is blocked off, no pressure was created on the output stroke so that suction on the suction stroke quickly brings a vacuum condition in the pumping chamber  106 . The diaphragm  34  and plunger  42  stay too far left and close port  121  and compress somewhat bias spring  126 .  
         [0055]    With reference to FIG. 9( b ), a configuration at mid-stroke is shown. The lower pressure in pumping chamber  106  which then causes a lower pressure in transfer chamber  44  holds diaphragm  34  and plunger  42  to the left but cannot hold them as far left as in the conventional pump as shown in FIG. 4( b ), because of the stiff bias spring with the higher spring constant  146 . Overfill of transfer chamber  44  is consequently limited to the volume of stretch of diaphragm  34  under these conditions.  
         [0056]    The suction stroke reaches its end in FIG. 9( c ) at bottom dead center. The high suction in the pumping chamber is still present, but the stiff spring (see reference point  2  in FIG. 10) counterbalances the suction force thereby raising the pressure in transfer chamber  44  and preventing overfilling of transfer chamber  44  prior to starting the output stroke. For example, in a preferred case, the differential pressure in the transfer chamber versus the pumping chamber is about 10.5 psi for the bias spring to counterbalance.  
         [0057]    The output stroke begins as shown in FIG. 9( d ). Piston  46  moves to the left since there is very low pressure in the pumping chamber. Pressure does not build in the transfer chamber except as caused by the stiff bias spring  126 , so diaphragm  34 , plunger  42 , and piston  46  move together.  
         [0058]    At mid-stroke as shown in FIG. 9( e ), check valves  102  stay closed and the stiff spring  126  biases to cause leakage out of the transfer chamber rather than into it.  
         [0059]    The output stroke finishes as shown in FIG. 9( f ). Since transfer chamber  44  has not overfilled, diaphragm  34  does not balloon and normal operation continues in spite of the plugged inlet line to the pumping chamber. Hence, the stiff bias spring  126  prevents the failure mode described with respect to FIGS.  4 ( a )- 4 ( f ).  
         [0060]    Thus, once the valve spool moves past the shut off port, the stiff bias spring prevents it from moving much further. As shown in FIG. 10, at the normal port shutoff position (reference point  1 ), both the weaker spring and the stiffer spring have a force of just over  4  pounds, or about 3.5-4.5 psi pressure on the diaphragm. Thus, the positive oil drive bias of U.S. Pat. No. 3,775,030 is maintained. Now, however, as travel is continued towards the maximum spring compression, the stiff spring has over 12 pounds of force versus only about 5 pounds of force for the weak spring. The added force limits the ability of the diaphragm to move too far under high vacuum conditions. This is true because the pull from the oil transfer chamber side is now the spring force plus the pressure differential between the pumping chamber and the transfer chamber. The conventional weak spring could only effectively counteract about 5 psi of vacuum; the improved stiff spring is optimized at counteracting about 10.5 psi of vacuum, which is all that is practically attainable (although theoretically, 14.7 psi could be obtained). Although designing for the highest force possible would assure that oil never is pushed into a full transfer chamber, it is only necessary that there is not a net increase in oil during a full suction and output cycle of the pump. In other words, as long as there is more time during the suction and output strokes where the hydraulic transfer chamber is above atmospheric pressure than below, there will be no average increase of oil in the chamber.  
         [0061]    Vacuum diaphragm rupture testing was done. Test results are shown in Table 1. A pump as described in FIG. 2 was used modified to have stiffer spring constants for bias spring  126  as shown in Table 1. A vacuum was maintained at the inlet (check valve  36 ). The vacuum was maintained at 15 in. Hg or less for a few hours and then was increased to 20 in. Hg or greater until failure or until the test was stopped.  
                                                           TABLE 1                       Test   Ser. No.   R   Run Time   Outcome                   1   141849   43.1 lb/in    97 hr   Rupture       2   141849   43.1    55   Rupture            Comment: Burr found; valve housing interior deburred            3   141849   43.1   106   Rupture       4   142132   53.7   106   OK       5   ?   53.7   124   OK       6   142131   53.7   214   OK                  
 
         [0062]    The first three tests were run with a stiff spring having a spring constant of 43.1 lb/in. The diaphragm ruptured at 97 hr. during the first test and at 55 hr. during the second test. After the second test, the pump was examined and a burr was found in the valve housing so that valve spool  84  was sticking so that eventually the diaphragm ballooned and got caught on base plate  90 . The valve housing was deburred and test 3 was run. The diaphragm ruptured at 106 hr. It was determined that the burr was not material to the findings except for time to failure. The 43.1 lb/in rated spring allowed failure to occur at about 100 hours.  
         [0063]    Tests 4-6 were run using a bias spring having a spring constant of 53.7 lb/in. In each test, the pump ran for over 100 hr. and for Test 6, the pump ran for over 200 hr. without diaphragm rupture.  
         [0064]    It was determined from the testing that the bias spring having the spring constant of 43.1 lb/in. was marginally acceptable. Clearly the pump having the bias spring with spring constant 53.7 lb/in. was acceptable since there were no failures. The conclusions of the testing are shown in FIG. 11. Line  150  shows the bias spring having spring constant of 43.1 lb/in. Line  148  shows the bias spring having spring constant of 53.7 lb/in. Broken line  152  represents a bias spring having a spring constant which would be the maximum ever needed. That is, the maximum vacuum which could be achieved at reference point  2 , the point at which base plate  88  contacts wall  108  (see FIG. 4( e )) is 14.7 psia. A pump like this could never achieve such a vacuum. Therefore, line  152  is shown as being broken and somewhat approximate. In any case, it gives the general idea of where a maximum spring constant would be.  
         [0065]    For a particular pump, the spring constant can be calculated in the following way assuming the following design assumptions. First, the diaphragm&#39;s equivalent area at mid-stroke is approximately the same as the piston area. Second, the minimum pressure differential across the diaphragm needed must be equal to the suction pressure the pump is designed for. Third, the maximum pressure differential is 14.7 psi. Based on that, the following statements can be made:  
         [0066]    1. Overfill distance is the difference in distance between the diaphragm and the piston at (i) maximum overfill position and (ii) neutral position (valve just closed).  
         [0067]    2. Overfill spring force is design suction pressure differential times the piston area.  
         [0068]    3. Neutral spring force is the neutral operating pressure differential times the piston area.  
         [0069]    4. Spring constant is the quantity of overfill spring force minus neutral spring force divided by the overfill distance.  
         [0070]    Based on these assumptions and statements, spring constant can be calculated from:  
           k=A   p ( P   s   −P   n )/ d   0    
         [0071]    where k is spring constant,  
         [0072]    A p  is piston area,  
         [0073]    d 0  is overfill distance,  
         [0074]    P s  is design suction pressure differential,  
         [0075]    P n  is neutral operating pressure differential.  
         [0076]    Based on the testing discussed above, appropriate maximum design suction pressure differential is 8.4-14.7 psia. Appropriate neutral operating pressure differential is zero to 4 psia.  
         [0077]    It is noted from FIGS. 10 and 11 that the stiffer bias spring of the present invention is necessarily shorter than the conventional spring. This has a good benefit in that when the pump is shut-down, the bias spring does not continually force oil out of the transfer chamber and past the piston assembly/housing interface to the reservoir. With the stiffer spring, once the transfer chamber has properly filled and the pump is turned-off, the spring no longer exerts a significant force. That means the transfer chamber has an oil fill which is at its proper pumping point, and it does not have to refill at the next start-up. On the other hand, the shorter spring does create a negative. The shorter spring does not fully expel air from the transfer chamber prior to initial start-up. The added air makes it very difficult to fully prime the transfer chamber  44 . In this case, the pump must be taken apart and manually primed or vacuum-primed for each of the several transfer chambers. Furthermore, sometimes the pump loses prime under conditions where air in the oil can accumulate and not be expelled. To address these negatives, notch  116  was developed. Notch  116  is a mechanism for expelling air. FIGS.  12 ( a )- 12 ( f ) show the operation of a pump having notch  116  with respect to bleeding air off and providing the further benefit of allowing the pump to self-prime.  
         [0078]    In FIG. 12( a ), the suction stroke begins. Transfer chamber  44  has an excess of air. Oil flows through open valve port  98  and pushes air to the high point in cylinder  47 . As the suction stroke starts, more oil wants to enter through check valves  32  and valve port  98 , but stiff bias spring  126  holds diaphragm  32  to move along with piston  46 .  
         [0079]    At mid-stroke as shown in FIG. 12( b ), there is a higher suction so that diaphragm  32  is pulled to the left to shut off valve port  121 . The stiff bias spring  126  resists compressing excessively so that diaphragm  32  moves substantially with piston  46 .  
         [0080]    As shown in FIG. 12( c ), there is still a high suction in the pumping chamber  106  as piston  46  nears its end stroke (BDC). The stiff spring limits the diaphragm plunger  42  and diaphragm  34  from going too far left and raises the pressure in the transfer chamber  44  to prevent oil overfill.  
         [0081]    As the output stroke begins as shown in FIG. 12( d ), piston  46  starts moving to the left, while check valves  32  close, and pressure in transfer chamber  44  builds. The rising pressure in transfer chamber  44  pushes air out notch  116 .  
         [0082]    At mid-stroke as in FIG. 12(e), pressure in transfer chamber  44  is above the reservoir pressure, and air continues to be pushed through notch  116 .  
         [0083]    At the end of the output stroke as in FIG. 12( f ), diaphragm  34  moves left as piston  46  moves left. Most of the air in transfer chamber  44  has now been expelled. As subsequent suction and output strokes proceed, all of the air gets expelled and the pump rapidly self-primes itself.  
         [0084]    Notch  116  can be square, hemispherical, triangular, or any shape. Notch  116  must be large enough to allow air to rather rapidly bleed off, but not so large that pump efficiency will suffer. Generally, a 1% loss of pump efficiency is acceptable. For a particular pump, it is then necessary to calculate an equivalent cross-sectional area for notch  116  which would be equivalent to the 1% loss of efficiency.  
         [0085]    As indicated earlier, the notch  116  should be placed at the top of the cylinder  120  so that it is located at the point where air would collect. The notch  116  should be long enough so that it is exposed to the pressurized oil zone for at least part of the piston stroke. It may extend to the end of the piston travel so that it is exposed for the entire stroke. The best practice is to have it exposed for the first half of the stroke only. The notch size must be large enough to allow rapid passage of air, and small enough to resist oil passage so that pump performance is not significantly reduced.  
         [0086]    For most pumps the cross sectional area of the notch  116  should be about 0.0002 square inches and height of 0.017 inches. To purge air effectively the cross sectional are should be greater than 0.00005 square inches. The maximum cross sectional area would be about 0.003 square inches. The height and width of the groove cross-section should both be greater than 0.005 inches.  
         [0087]    The improved pump of the present invention results in improved reliability because premature diaphragm ruptures caused by unintended hydraulic oil over-fill of the transfer chamber is eliminated. The improved pump results in improved efficiency and smoothness of output because the fully intended diaphragm stroke length is continually utilized because there is less air left in the transfer chamber during normal operation. The pump of the present invention has an improved metering capability of oil/air relative to the transfer chamber and reservoir thereby ensuring a consistently high quality of oil within the transfer chamber and thereby maintaining the “stiffest” hydraulic system practical, regardless of pump inlet and outlet conditions. The pump of the present invention self-primes and avoids any loss of prime during operation. Thus, the pump of the present invention is significantly improved over the conventional diaphragm pump.  
         [0088]    The above specification, examples and data provide a complete description of the manufacture and use of the composition of the invention. Since many embodiments of the invention can be made without departing from the spirit and scope of the invention, the invention resides in the claims hereinafter appended.