Abstract:
A control valve is fluidly coupled to chambers containing fluid of different pressures for regulating flow therebetween. The control valve has a valve housing having a chamber fluidly coupled to a first chamber, a second chamber, and a third chamber. A fluid flow regulation member is disposed in the chamber and is configured to regulate fluid flow between the second chamber and the third chamber. A diaphragm is disposed substantially perpendicular to a longitudinal axis of the chamber in which longitudinal deflection of diaphragm is representative of the pressure in the first chamber.

Description:
In accordance with the provisions of 35 U.S.C. 119(e), Applicant hereby claims the priority of U.S. Provisional Patent Application No. 60/260,357, filed Jan. 8, 2001. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to a control valve, and more particularly, to a control valve for a variable displacement compressor, such as commonly used in air conditioning systems. 
     DESCRIPTION OF RELATED ART 
     FIG. 8 schematically depicts an air conditioning system, such as that used in an automobile to provide passengers a comfortable atmosphere. Air conditioning systems typically include a compressor  100 , a condenser  102 , an expansion device  104 , and an evaporator  106  fluidly connected together by tubes or hoses  108  in which refrigerant flows. In order to condition the air before it is released to the passenger compartment, heat is removed from the air by passing the air through the evaporator  106 . This causes the refrigerant to boil and form a gas, which travels from the evaporator  106  to the compressor  100 . The compressor  100  serves as a pump for circulating the refrigerant through the entire system. In addition, the compressor  100  may increase the temperature and pressure of the refrigerant. 
     Vehicle air conditioning systems commonly use variable displacement compressors, which allow the adjustment of the refrigerant pumping capacity in response to the air conditioning load. The compressor  100  comprises three main chambers, which include a suction chamber  110 , a crankcase chamber  112 , and a discharge chamber  114  with a valve plate  116  separating the three chambers. This valve plate  116  contains ports fluidly coupling the suction chamber  110  to other areas of the compressor  100 . 
     Refrigerant flowing from the evaporator  106  enters the compressor  100  through the suction chamber  110  located in the rear head  118  of the compressor  100 . The refrigerant flows into the suction chamber  110  into a cylinder  122  through a port  120  where pistons  124  compress the refrigerant. The compressed refrigerant exits through discharge port  126  into the discharge chamber  128  coupled to the condenser  102  by a tube or hose  108 . The pressure of the refrigerant in the discharge chamber  114  always exceeds both the pressure of the refrigerant in the suction chamber  110  as well as the crankcase chamber  112 . 
     The pumping capacity of the pistons  124  may be adjusted by changing the inclination angle θ of a swashplate  130  relative to the compressor shaft  132 . The pumping capacity corresponds to the stroke length of the piston  124 . A larger stroke length corresponds to a higher pumping capacity and a higher pressure in the discharge chamber  114 . Similarly, a lessening stroke length corresponds to a decreased pumping capacity and a lower pressure in the discharge chamber  114 . The inclination angle θ of the swashplate  130  relates directly to the piston  124  stroke length. 
     The swashplate  130  is located in the crankcase chamber  112  and is connected by pivot  134  to the compressor shaft  132  and the pistons  124 . The angle formed between the connection point of the swashplate  130  and the rotation of the swashplate  130  represents the inclination angle θ. The rotational movement of the compressor shaft  132  rotates the swashplate  130  causing the pistons  124  to reciprocate in their cylinders.  122 . The compressor shaft  132  moves responsive to the vehicle engine via a pulley  136  with the compressor shaft  132  being mounted on radial bearings  138  and shoes  140 , which allows the swashplate  130  to rotate. 
     The crankcase chamber  112  contains refrigerant leaked by the pistons  124 . Variable displacement of the compressor  100  is obtained by varying the crankcase chamber  112  pressure Pc relative to the suction chamber  110  pressure Ps. Changing the pressure differential (Pc−Ps) between the crankcase chamber  112  and the suction chamber  110  causes the inclination angle θ of the swashplate  130  to vary, which regulates the pumping capacity of the pistons  124 . 
     A small pressure differential (Pc−Ps) corresponds to an increased inclination angle θ. When the inclination angle θ is at its maximum, the pistons  124  reciprocate at the maximum stroke thus highest compression. At this point, the air conditioning system is at its highest cooling capacity. In contrast, an increasing pressure differential (Pc−Ps) corresponds to a decreasing inclination angle θ. Decreasing the inclination angle θ causes the pistons  124  to de-stroke resulting in lower compression. At this point, the air conditioning system is at its lowest cooling capacity. 
     For example, if the pressure differential Pc−Ps is low, such as 5-15 kPa, the compressor operates at maximum stroke with the swashplate  130  at its maximum inclination angle θ. In contrast, if the pressure differential Pc−Ps is high, such as 100-150 kPa, the compressor operates at minimum stroke with the swashplate  130  at its minimum inclination angle θ. At this point, the swashplate  130  is nearly perpendicular to the compressor shaft  130 . A de-stroke spring  131  in FIG. 8 is provided to force the swashplate  130  to this position when cooling capacity is not needed. 
     Reference is made to U.S. Pat. No. 6,146,106 illustrating a control valve consistent with the prior art. FIG. 9 schematically illustrates the control valve  144  of the &#39;106 patent which may be used with the compressor schematically illustrated in FIG.  8 . The variable displacement compressor  100  uses a control valve  144  to regulate the pressure differential (Pc−Ps). The suction chamber  110  pressure Ps changes as certain parameters in the car change, such as compressor speed. This has a direct effect on the pressure differential (Pc−Ps). The control valve  144  adjusts the pressure Pc in the crankcase chamber  112  relative to the pressure Ps in the suction chamber  110  in order to reach an equilibrium point. The equilibrium point is the set pressure differential (Pc−Ps) value of the control valve. By maintaining a constant pressure differential (equilibrium point), the cooling air entering the passenger compartment stays relatively constant regardless of changing parameters. 
     The control valve  144  regulates the flow of refrigerant from the discharge chamber  114  having a discharge chamber pressure Pd to the crankcase chamber  112  relative to the pressure of the refrigerant in the suction chamber  110 . The control valve  144  contains a bellows  146 , which compresses or expands as a result of an increase or decrease, respectively, of the fluid in the suction chamber  110 . When there is a high pressure differential Pc−Ps, the control valve  144  allows more refrigerant to flow from the discharge chamber  114  into the crankcase chamber  112  than can escape to the suction chamber  110  through flow passage  148 . The flow passage  148  is sized so that the amount of flow from crankcase chamber  112  to suction chamber  110  is less than the flow from the discharge chamber  114  to the crankcase chamber  112 . As a result, the crankcase chamber pressure Pc increases, causing the compressor  100  to de-stroke. When the compressor  100  de-strokes, the suction chamber pressure Ps increases as a result of reduced refrigerant flow out of the compressor  100 . The bellows  146  of the control valve  144  responds accordingly, reducing the flow into the crankcase chamber  112  until equilibrium is reached. 
     The bellow  146  connects to a poppet  150  or other type of member for regulating the flow from the discharge chamber  114  to the crankcase chamber  112 . When the compressor  100  begins to de-stroke as the result of a high-pressure differential, the suction chamber  110  pressure increases. The fluid from the suction chamber  110  acts on the exterior of the bellows  146 . An increasing suction chamber  110  pressure causes the bellows  146  to decrease in length. This moves the poppet  150  in a direction to reduce the flow from the discharge chamber  114  to the crankcase chamber  112  until the poppet  150  rests at the equilibrium point. Traditionally, the equilibrium point had a fixed setting, i.e. a set pressure differential between the crankcase chamber  112  and the suction chamber  110 . 
     With the development of improved air conditioning systems and an increased emphasis on fuel economy, it was desired to vary the equilibrium point for a closer matching of compressor capacity to load. Solenoid-actuated control valves provide one means for varying the equilibrium point. The solenoid-actuator  152  connects to the poppet  150 , which regulates fluid flow between the discharge and crankcase chamber  114 ,  112 . As such, the solenoid actuator  152  may vary the fluid flow regardless of the pressure from the suction chamber  110 . This in turn varies the equilibrium point. An electrical controller  154  connects to the solenoid for varying the amount of current supplied to the solenoid. The amount of supplied current may be set in response to various parameters, such as engine speed, vehicle speed, cabin air temperature, etc. This in turn moves the poppet  150  to a different equilibrium point. 
     The resultant design incorporated a mechanical bellow control valve with an electrical solenoid-actuator. This design, however, presents certain concerns. Compressors in vehicles must operate in a wide range of conditions. These conditions range from extreme heat to extreme cold. Moreover, compressors experience significant amounts of vibration from the road, vibration of the engine, etc. As a result, the bellows undergoes significant amounts of wear and tear, which reduces the bellows&#39; useful life. As the bellows are relatively long, the vibrations cause the bellows to vibrate and contact the internal surfaces of the control valve. Over time, the bellows have been observed to break down and lose their resiliency resulting in a less efficient air conditioning system. Once a bellows fails, typically the complete control valve must be replaced in order for the air conditioning system to work properly. However, bellows require a significant manufacturing process, increasing their replacement cost. 
     Accordingly, a need exists to increase the useful life of a vehicle air conditioning system, and for a control valve that will better resist the hostile environment conditions experienced in a vehicle compressor. 
     SUMMARY OF THE INVENTION 
     These and other needs are met by the present invention, which provides a variable displacement compressor having a suction chamber, a crankcase chamber, and a discharge chamber. The crankcase chamber and discharge chamber are fluidly coupled by a valve for regulating the flow therebetween as a function of pressure in the suction chamber. The valve comprises a valve housing having a chamber fluidly coupled to the suction chamber, the crankcase chamber, and the discharge chamber. The fluid flow regulation member is disposed in the chamber and is configured to regulate fluid flow between the crankcase chamber and the discharge chamber. A diaphragm is disposed substantially perpendicular to a longitudinal axis of the chamber and acts on the fluid flow regulation member as a function of the pressure in the suction chamber, the amount of longitudinal deflection of the diaphragm being responsive to the pressure in the suction chamber. 
     The control valve may be applied to other applications requiring the regulation of flow between two chambers relative to another chamber. This control valve is fluidly coupled to chambers containing fluid of different pressures for regulating flow therebetween. The control valve comprises a valve housing having a chamber fluidly coupled to a first chamber, a second chamber, and a third chamber. A fluid flow regulation member is disposed in the chamber and is configured to regulate fluid flow between the second chamber and the third chamber. A diaphragm disposed substantially perpendicular to a longitudinal axis of the chamber in which longitudinal deflection of diaphragm is representative of the pressure in the first chamber. 
     The deflection of the diaphragm discussed above acts on the fluid flow regulation member. The diaphragm has an outer perimeter shape substantially corresponding to the shape of the chamber perpendicular to the longitudinal axis. The diaphragm is configured to deflect in a first axial direction as a function of increasing force acting on the diaphragm and deflect in a second axial direction as a function of decreasing force acting on the diaphragm. In contrast, embodiments of the invention, the diaphragm comprises an undulation having at least one ridge and at least one groove. This undulation of the diaphragm compresses or expands along the axis perpendicular to the longitudinal axis of the chamber with the longitudinal deflection of the diaphragm. The outer periphery of the diaphragm is hermetically sealed to the inner wall of the chamber creating a volume between the diaphragm and an end of the chamber. A vacuum exists in this volume. 
    
    
     The foregoing and other features, aspects, and advantages of the present invention will become more apparent from the following detailed description of the present invention when taken in conjunction with the accompanying drawings. 
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 a  depicts a cross-sectional view of the control valve of the present invention. 
     FIG. 1 b  depicts a cross-sectional view of the valve housing of the control valve illustrated in FIG. 1 a.    
     FIG. 2 depicts an oblique view of one embodiment of the diaphragm. 
     FIG. 3 a  depicts the diaphragm of FIG. 2 in a state of no deflection. 
     FIG. 3 b  depicts the diaphragm of FIG. 2 in a first deflection state. 
     FIG. 3 c  depicts the diaphragm of FIG. 2 in a second deflection state. 
     FIG. 4 depicts a diaphragm having a high undulation frequency. 
     FIG. 5 a  depicts a substantially planar diaphragm in accordance with certain embodiments of the invention. 
     FIG. 5 b  depicts a convex shaped diaphragm in accordance with embodiments of the invention. 
     FIG. 6 a  schematically illustrates a valve housing with a diaphragm in a state of no deflection. 
     FIG. 6 b  schematically illustrates the valve housing with a diaphragm in a first state of deflection. 
     FIG. 6 c  schematically illustrates the valve housing with a diaphragm in a maximum deflection state. 
     FIG. 7 a  schematically illustrates the valve housing connected to a variable displacement compressor with a diaphragm in a state of deflection. 
     FIG. 7 b  schematically illustrates the valve housing connected to a variable displacement compressor with a diaphragm in a state of no deflection. 
     FIG. 8 schematically depicts an air conditioning system using a variable displacement compressor of the prior art, in which the control valve of the present invention can be employed. 
     FIG. 9 schematically illustrates a control valve of the prior art incorporating a bellows. 
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     The present invention addresses and solves problems associated with the degradation of control valves and more particularly to control valves in variable displacement compressor systems. A diaphragm is provided to form a pressure control member, which increases the useful life of the control valve. 
     FIGS. 1 a  &amp;  1   b  depict a cross-sectional view of the control valve  2  and the valve housing  4  of the control valve  2 , respectively, of the present invention. This control valve  2  may be incorporated into a variable displacement compressor  100  of the prior art, such as that shown in FIG.  8 . However, the control valve  2  may also be used with other applications requiring a control valve  2  responsive to pressure differentials. 
     The control valve  2  comprises a valve housing  4  having an inner chamber  5 . The valve housing  4  comprises a valve body  6  that is substantially cylindrical and a housing cap  10 . A first chamber  12  is fluidly coupled to the inner cavity  8  via a first fluid port  14  through the valve body  6 ; a second chamber  16  via a second fluid port  18 ; and a third chamber  20  via a third fluid port  22 . In this inner cavity  8 , the control valve  2  controls fluid flow between the second chamber  16  and third chamber  20  as a function of the pressure in the first chamber  12 . 
     A pressure control member  24  and a fluid flow regulation member  26  are both disposed in the chamber  5  of the valve housing  4 . The pressure control member  24  comprises a diaphragm  24 , which deflects a longitudinal direction (axial) direction with changes in fluid pressure from the first chamber  12 . The diaphragm  24  replaces the bellows traditionally used in control valves for controlling pressure. The diaphragm  24  controllably deflects as a function of changing pressure of the fluid received from the first chamber  12 , as a bellows  146  is designed to do. However, the diaphragm  24  corrects certain problems associated with a bellows  146 . The described diaphragm  24  assembly occupies significantly less volume as compared to the bellows design described by the prior art. As a result, the overall control valve  2  may be made much smaller compared to conventional designs. Since the diaphragm  24  is constructed of a rigid material and occupies significantly less volume, the diaphragm  24  resists the vibrations common in control valve  2  applications and therefore does not rub against opposing surfaces. As a result, the diaphragm  24  does not experience the same wear, as does a traditional bellows design. 
     The diaphragm  24  is contained by the valve housing  4  in a cavity  28  separate from the inner cavity  8 . The housing cap  10  mounts on one end of the valve body  6  forming the valve housing  4 . The housing cap  10  forms a cavity  28  wherein the diaphragm  24  is mounted. Both the inner cavity  8  and cavity  28  form the inner chamber  5  of the valve housing  4 . The housing cap  10  may be press-fit to the valve body  6  or secured by other suitable means. The diaphragm  24  is hermetically sealed to a flange  30  in the inner wall of the housing cap  10 . This creates a volume  32  between the diaphragm  24  and the underside surface  34  of the housing cap  10 . It is preferable for a substantial vacuum to exist in this volume  32 . Absent a vacuum, the diaphragm  24  would have very limited deflection characteristics. To create a vacuum, the diaphragm  24  is hermetically sealed to the flange  30  housing cap  10  under vacuum conditions. Once hermetically sealed and removed from the vacuum conditions, the volume  32  retains the vacuum applied during assembly. The diaphragm  24  may be hermetically sealed by electron beam welding, laser welding, pressing and retaining an O-ring, brazing, or other suitable means to create a hermetic seal. 
     As another alternative, the volume  32  may be filled with a gas having an expansion rate different from the expansion rate of the fluid received from the first chamber  12 . The expansion rate of the gas and fluid typically correspond to a change in temperature of the gas or fluid The selection of gas allows a designer to control the deflection characteristics of the diaphragm and the overall operating characteristics of the system to which the control valve  2  is applied. 
     A pin  36  is provided with one end interacting with the diaphragm  24  and the other end communicating with the fluid flow regulation member  26 . A stop member  38  attached at one end of the inner cavity  8  secures one end of the spool spring  40 , and provides a guide for the pin  36 . The stop member  38  has a central aperture  42  in which the pin  36  may reciprocate. The central aperture is aligned with an aperture in the valve body  6  such that the pin  36  is constrained to move axially through both apertures and contact the diaphragm  24 . Fluid from the first chamber  12  enters the inner cavity  8  through the first fluid port  14  and acts on the pin  36 . An increase in pressure of the fluid from the first chamber  12  correlates to an increased force acting on the pin  36 . This causes the diaphragm  24  to deflect in a first axial direction  44 . The pin  36  moves in the first axial direction  44  against the diaphragm  24  by an amount equal to the diaphragm  24  deflection. A decrease in pressure of the fluid from the first chamber  12  correlates to a decreased force acting on the pin  36 . Since the diaphragm  24  tends to return to its original shape, the diaphragm  24  forces the pin  36  in a second axial direction  46  as a result of the decreased pressure. The pin  36  thereby communicates the movement of the diaphragm  24  in either direction to the fluid flow regulation member  26 , discussed below. 
     In certain embodiments, the diaphragm  24  may be made of a rigid material, such as stainless steel, a Kapton polymer, and the like. The diaphragm  24  may be stamped from a sheet of material to form the desired shape. A designer should assess the common operating pressure range of the first chamber  12  and select a diaphragm  24  material and shape accordingly. It is preferred to select a material with a rigidity and shape such that at minimum pressure, the diaphragm  24  is in its original form, and at maximum pressure, the diaphragm  24  deflects to a maximum deflection position. Moreover, a material should be chosen that will resist any caustic effects of the fluid from the first chamber  12  if fluid were to leak into cavity  28 . 
     FIG. 2 depicts an oblique view of an exemplary diaphragm shape. FIGS. 3 a-c  schematically illustrate the diaphragm of FIG. 2 in a first, second, and third deflection state, respectively. As illustrated, the diaphragm  24  is disc-shaped and is corrugated to form an undulation from the diaphragm&#39;s  24  outer periphery  48  towards the center  50 . The undulation, which introduces yield or give into the diaphragm  24 , is a series of small ridges  52  and grooves  54  terminating towards the center  50  of the diaphragm  24 . As illustrated in FIG. 3 a , the diaphragm  24  shape has center portion  50  extending perpendicularly outwards with respect to a reference plane in line with the outer periphery  48  of the diaphragm  24  at a predetermined distance “a”. Increasing the frequency of undulation, illustrated by FIG. 4, introduces a greater yield into the diaphragm  24 . In other words, less force is required to deflect the diaphragm  24  with a higher frequency of undulation. A designer should consider the maximum pressure and force exerted by the fluid from the first chamber  12  in designing the diaphragm  24 . The diaphragm  24  should have a rigidity and an undulation frequency such that the diaphragm  24  deflects to a maximum position when the fluid from the first chamber  12  is at a maximum pressure and a maximum force is applied against the diaphragm  24  by a solenoid actuator  68  discussed below. 
     The following provides a description of the forces acting on the diaphragm  24  as the result of an axial force applied at the center  50  of the diaphragm  24 . When the diaphragm  24  deflects in the first axial direction  44 , as illustrated by FIG. 3 b , each ridge  52  and groove  54  moves closer to an adjacent ridge  52  and groove  54 , respectively. In other words, the undulation portion of the diaphragm  24  compresses in the direction perpendicular to the axial deflection, the horizontal direction. The amount of compression corresponds to the predetermined distance “a”. When the center  50  of the diaphragm  24  is in the same plane as its periphery  48  (distance “a”=0) as illustrated by FIG. 3 b , the diaphragm  24  is at a maximum compression. Compression arrows  56  illustrate the compression force acting on the diaphragm  24 . As illustrated by FIG. 3 c , when the center  50  of the diaphragm  24  moves in the first axial direction  44  past a reference plane in line with the outer periphery  48  of the diaphragm  24  by a distance “b”, the undulation portion of the diaphragm  24  expands, i.e. the ridges  52  and grooves  54  move away from adjacent ridges  52  and grooves  54 , respectively. The expansion forces acting on the diaphragm  24  are illustrated by expansion arrows  58 . At maximum deflection distance “c”, the diaphragm  24  contacts a stop surface  60  on the underside surface  34  of the housing cap  10 . 
     The total deflection of the diaphragm equals distance “a” plus distance “c”. When determining the frequency of undulation, the designer should consider the corresponding deflection distances “a” and “c”. For example, assume that one control valve  2  design requires the distance “a” be one distance unit and a second design requires the distance “a” be two distance units. For each design to function within the same pressure range, the second design would require a higher frequency of undulation than the first design to account for the increased total deflection. 
     The shape of the diaphragm  24  correlates to the force required to deflect the diaphragm  24  in the first axial direction  44 . For example, if the diaphragm  24  is flat absent an undulation portion as illustrated in FIG. 5 a , the force required to deflect the diaphragm would be considerable, as the diaphragm  24  material must expand. The considerable amount of force required is a result of little yield in the diaphragm  24  due to the absence of an undulation portion to account for the expansion forces  58  acting on the diaphragm  24 . If the diaphragm  24  is convex in shape and absent an undulation as illustrated in FIG. 5 b , a substantial force would also be required to deflect the diaphragm  24  in the first axial direction  44 . The diaphragm  24  again contains little yield due to the absence of an undulation portion to accommodate the compression forces  56 . 
     The pin  36  (seen in FIGS. 1 a  and  1   b ) communicates the axial deflection of the diaphragm  24  to the fluid flow regulation member  26 . As illustrated in FIGS. 1 a  &amp;  1   b , the fluid flow regulation member is provided by a spool  26  disposed in the inner cavity  8  of the valve body  6 . The spool  26  is cylindrical with a diameter corresponding to inner diameter of the inner cavity  8 . The spool  26  has a groove  62  around its outer periphery, which spans the second fluid port  18  and the third fluid port  22 . The volume created by the groove  62  and the wall of the inner cavity  8  contains fluid from the second and third chambers  16 ,  20  within the volume. Fluid from the first chamber  12  introduced into the inner cavity  8  through the first fluid port  14  is prevented from interacting with the fluid from the second and third chamber  16 ,  20  by the spool  26 . 
     The spool  26  reciprocates within the inner cavity  8  responsive to a force applied in the first axial direction  44  by a solenoid actuator  68  and a force applied in the second axial direction  46  by the diaphragm  24  via the pin  36  and spool spring  40 . The functions of the solenoid actuator  68  and the spool spring  40  are discussed below. The opposing forces acting on the spool  26  regulate the rate of fluid flow between the second chamber  16  and the third chamber  20  as a function of the pressure in the first chamber  12 . When the spool  26  moves axially, the edge  64  of the groove  62  passes over the third fluid port  22 . Depending on the direction of movement, the third fluid port  22  is either increasingly or decreasingly closed to regulate the fluid flow rate. 
     In addition to the pin  36  movement, the spool spring  40  biases the spool  26  in the second axial direction  46 . In one embodiment, the spool spring  40  is coiled around the pin  36  but is not physically attached to the pin  36 . One end of the spool spring  40  rests on the stop member  38  with the spring circumference surrounding the aperture  42  through which the pin  36  passes. This allows both the pin  36  and the spool spring  40  to move freely with respect to one another. The pin  36  does, however, keep the spool spring  40  from buckling to one side when the spool spring  40  is compressed. The spool spring  40  may also be provided separate from the pin  36 . However, for the above reasons, it is preferred that the spool spring  40  is coiled around the pin  36 . 
     The fluid from the first chamber  12  enters the inner cavity  8  through first fluid port  14  positioned below the pin  36  and diaphragm  24 . An increase in fluid pressure from the first chamber  12  forces the pin  36  in the first axial direction  44 . However, in order to avoid the same force acting on the spool  26  which would force the spool  26  in the second axial direction  46 , fluid from the first chamber  12  flows through a spool aperture  66  to the opposite end of the spool  26 . As a result, the pressure of the fluid acts equally on each end of the spool  26 ; changes in the pressure of the first chamber  12 , therefore, have no direct effect on the spool  26  movement. The only effect of the pressure is that communicated by the diaphragm  24  via pin  36 . 
     As illustrated by FIG. 1 a , the solenoid-actuator  68  of the control valve  2  generates an opposing force acting on the spool  26  against the diaphragm  24  and spool spring  40 . The diaphragm  24  provides feedback to maintain the pressure differential point for a given applied solenoid current. The solenoid actuator  68  comprises an armature spring  70 , an armature  72 , and a rod  74  contained in an armature housing  76 . A coil housing  78  enclosing a coil  80 , which carries a magnetic flux, surrounds the armature housing  76 . Current applied to the coil  80  creates a magnetic flux generated by coil  80  acting on the armature  72  attracting it towards the pole surface  87 . The armature  72  and rod  74  move in the second axial direction  44 . A flux ring  82  disposed between the coil housing  78  and armature housing  76  directs the magnetic flux to the armature spring  70  and armature  72 . Retaining clips  84  secure the coil housing  78  to the armature housing  76 . 
     In order to integrate the solenoid actuator  68  with the valve housing  4  and more particularly, the valve body  6 , a pole section  86  is mounted between the valve body  6  and the armature housing  76 . One end of the pole section  86  has an outer diameter consistent with the inner diameter of the armature housing  76 . The other end flanges radially outwards in which the valve body  6  is disposed. The valve body  6  and the armature housing  76  are hermetically sealed to the pole section  86  to prevent fluid leakage. The pole section  86  has a central aperture  88  in line with an aperture  90  of the valve body  6 . The rod  74  reciprocates within apertures  88 ,  90  with one end interacting with the spool  26  and the other end with the armature  72 . The armature spring  70  is disposed with one end on the armature housing  76  and the other end interacting with the armature  72 . 
     The spool spring  40  is of a length such that when no electric current is applied to the coil  80 , the diaphragm  24  in an undeflected state does not apply a force via pin  36  on the spool  26 . Also, the lengths and stiffness of the spool spring  40  and the armature spring  70  are chosen such that the groove  62  of the spool  26  spans both the second and third fluid ports  18 ,  22  when no electrical current is applied to the coil  80 . Additionally, the spool spring  40  forces the rod  74 , armature  72 , and armature spring  70  in the second axial direction  46 . 
     When the solenoid actuator  68  force applied to the spool  26  in the first axial direction  44  equals the force in the second axial direction  46 , the spool  26  does not move. At this point, there is a constant fluid flow between the second chamber  16  and the third chamber  20 . This point is also known as the equilibrium point. In other words, the equilibrium point is the point at which the force applied by the diaphragm  24  via pin  36  and spool spring  40  equals the opposing force applied by the solenoid actuator  68 . The equilibrium point also represents the corresponding pressure differential between the first and second chambers  12 ,  16 . An electric controller  92  connects to the solenoid actuator  68  to vary the current and thus the equilibrium point. 
     The solenoid actuator  68  may be replaced with a spring, diaphragm, or other type of resilient element to force the spool  26  in the first axial direction  44 . However, in this case, the control valve  2  would have a fixed equilibrium point, as one could not vary the applied force in the first axial direction  44 . Such configurations can be advantageous depending on the control valve&#39;s  2  application. 
     In the manufacture of the control valve  2 , the diaphragm  24  and spool spring  40  should be chosen to have deflection characteristics to correspond to the minimum current I( 1 ) and maximum current I( 2 ) applied to the coil  80  of the solenoid actuator  68 . At minimum current I( 1 ), the spool spring  40  should force the spool  26  to a position where fluid flow between the second and third chambers  16 ,  20  is maximized. At the second current I( 2 ), the solenoid actuator  68  should force the spool  26  to a position of minimum flow between the second and third chambers  16 ,  20 . At this point, both the spool spring  40  and diaphragm  24  will be at a maximum deflection. 
     As the diaphragm  24  allows the control valve  2  to be manufactured significantly smaller than prior art control valves, each element of the control valve  2  is preferably manufactured to greater precision. Therefore, the position of the spool  26  may need fine tuning after manufacture. For example, after assembly, if the flow rate at the applied current I( 1 ) or I( 2 ) does not meet specifications, the end  94  of the armature housing  76  may be deformed inwards. 
     This adjustment moves the armature spring  70 , armature  72 , rod  74 , and spool  26  in the first axial direction  44  thereby altering the fluid flow rate between the second and third chamber  16 ,  20 . 
     Referring to FIGS. 6 a-c , the following discusses the movement of the diaphragm  24 , pin  36 , spool  26 , and solenoid actuator  68 . In each of the figures, the solenoid actuator  68  is not shown. However, the position of the rod  74  illustrates the corresponding force, i.e. current, applied by the solenoid actuator  68 . 
     FIG. 6 a  schematically illustrates the diaphragm  24  in an undeflected state and the current I( 1 ) applied to the solenoid actuator  68 . The fluid from the first chamber  12  is at a pressure such that the diaphragm  24  does not deflect. Also, the current applied to the solenoid actuator  68  in FIG. 6 a  does not move the spool  26  in the second axial direction  46 . Therefore, as illustrated, the groove  62  of the spool  26  spans both the second fluid port  18  and the third fluid port  22 . The edge  64  of the groove  62  does not cover the third fluid port  22 , thereby allowing maximum fluid flow between the second and third fluid ports  18 ,  22 . The spool spring  40  is also in its maximum expanded position forcing the spool  26  and rod  74  to the furthest position in the second axial direction  46 . 
     If the current applied to the solenoid actuator  68  increases and/or the pressure of fluid from the first chamber  12  increases, the spool  26  moves in the second axial direction  46  as illustrated by FIG. 6 b . In the first case, if the electric controller  92  increases current to the solenoid actuator  68 , the control valve  2  elements are forced in the first axial direction  44  to decrease the fluid flow rate between the second and third chambers  16 ,  20 . In the second case, the increase in pressure of fluid from the first chamber  12  causes the diaphragm  24  to deflect in the first axial direction  44 . As a result, the force applied to the spool  26  by the diaphragm  24  via pin  36  decreases. If the solenoid actuator current is of a value to overcome the force applied by spool spring  40 , the solenoid actuator  68  forces the spool  26  in the first axial direction  44 , compressing the spool spring  40 . The spool  26  stops at the position where the forces applied in the first axial direction  44  equal the forces applied in the second axial direction  46 . The spool  26  may move to the illustrated position due to a combination of conditions described with respect to the first and second case as well. The edge  64  of the groove  62  partially covers the third fluid port  22  which decreases the fluid flow rate between the second and third chambers  16 ,  20 . 
     FIG. 6 c  illustrates the spool  26  and diaphragm  24  in the maximum state of deflection. As discussed with respect to FIG. 6 b , this may be a result of the current applied to the solenoid actuator  68 , increased pressure from the first chamber  12 , or a combination of both conditions. As illustrated, the distance of maximum diaphragm  24  deflection corresponds to a minimum flow rate between the second and third chamber  16 ,  20 . The edge  64  of the groove  62  completely covers the third fluid port  22 , thereby stopping the flow between the second and third chambers  16 ,  20 . This state may not be desirable as it could introduce an overpressure situation. The control valve  2  may be designed to allow some flow between the second and third chambers  16 ,  20  during this state. 
     The functions of the elements described above may be better understood with respect to the control valve  2  application in a variable displacement compressor  100 . 
     As discussed with respect to the prior art, a variable displacement compressor  100  comprises three main pressure chambers, which include the suction chamber  110 , the discharge chamber  114 , and the crankcase chamber  112 . The suction chamber  110  connects to the first fluid port  14 , the crankcase chamber  112  to the second fluid port  18 , and the discharge chamber  114  to the third fluid port  22 . The discharge chamber  114  contains refrigerant that is under high pressure. The fluid contained by the discharge chamber  114  is at a pressure greater than the fluid contained by either the suction chamber  110  or crankcase chamber  112 . Further, the fluid pressure of the crankcase chamber  112  is greater than the fluid pressure in the suction chamber  110 . Therefore, in order to increase the pressure in the crankcase chamber  112 , the control valve increases the flow from the discharge chamber  114  to the crankcase chamber  112 . 
     The equilibrium point is the pressure differential (Pc−Ps) between the crankcase chamber  112  and the suction chamber  110 . This equilibrium point also represents the point at which the force applied by the solenoid-actuator  68  equals the spool spring  40  and diaphragm  24  force. The suction pressure of the compressor  100  may increase due to a change in the system, such as an increase in thermal load on the evaporator. As illustrated by FIG. 7 a , the increased suction pressure causes the diaphragm  24  to deflect in the first axial direction  44 . The spool  26  moves in the same direction as a result of the force applied by the solenoid actuator  68 . This causes the flow from the discharge chamber  114  to the crankcase chamber  112  to decrease as the groove edge  64  covers a portion of the third fluid port  22 . Accordingly, the crankcase chamber  112  pressure decreases. Consequently, the pressure differential between the crankcase chamber and the suction chamber, Pc−Ps, also decreases. As a result, the pistons reciprocate at a higher stroke and thus higher compression and cooling capacity. The higher cooling capacity satisfies the increased thermal load on the evaporator. 
     Referring to FIG. 7 b , similarly, the suction pressure of the compressor  100  may decrease due to a decrease in required thermal load on the evaporator. Therefore, the diaphragm  24  deflects in the second axial direction  46  forcing the pin  36  against the spool  26 . The spool  26  also moves in this direction and increases the flow from the discharge chamber  114  to the crankcase chamber  112 , causing the pressure of the crankcase chamber  112  to increase. Consequently, the pressure differential Pc−Ps increases. As a result, the compressor de-strokes as a result of the lessening compression by the pistons. Therefore, the cooling capacity decreases as a result of the decreasing thermal load. 
     When minimum or no current is applied by the electrical controller  92  to coil  80 , the spool spring  40  forces the spool  26  to a position such that the groove  62  spans both the second and third fluid ports  18 ,  22 . The pressure differential at this point (Pc−Ps) is at a maximum value as the high pressure discharge fluid enters the crankcase chamber at a maximum rate. This corresponds to a minimum stroke condition and the least cooling capacity. 
     In order to increase the cooling capacity, an electrical controller  92  increases the applied current. The armature  72  is therefore forced in a first axial direction  44  by the magnetic force on the armature  72 . The rod  74  forces the spool  26  in the first axial direction  44  so as to decrease the fluid flow from the discharge chamber  114  to the crankcase chamber  112 . This position is also illustrated by FIG. 7 a . At this point, the spool  26  encounters the resistant force applied by the diaphragm  24  via pin  36  due to the pressure of the suction chamber  110 . A new equilibrium point is established where less fluid flows from the discharge chamber  114  to the crankcase chamber  112 . This corresponds to a higher stroke position and a higher cooling capacity. 
     For example, assume the solenoid actuator  68  applied force corresponds to an equilibrium point of 50 kPa. Further assume that the fluid pressure in the suction chamber  110  is 75 kPa; the crankcase chamber  112  has a pressure of 125 kPa; and the discharge chamber  114  has a pressure of 150 kPa. Therefore, the pressure differential between the crankcase chamber  112  and the suction chamber  110  (Pc−Ps) is 50 kPa, which is currently at equilibrium. If the pressure in the suction chamber  110  increase to 100 kPa, the pin  36  causes the diaphragm  24  to deflect in the first axial direction  44 . As a result, the solenoid actuator  68  forces the spool  26  in the first axial direction  44 . The spool  26  movement decreases the fluid flow between the crankcase chamber  112  and the discharge chamber  114 . As a result, the pressure of the fluid in the discharge chamber  114  will increase. The control valve  2  is designed to maintain the equilibrium point of 50 kPa. Therefore, the discharge chamber  114  pressure will increase to 175 kPa. Assume now that the fluid pressure in the suction chamber  110  drops to 50 kPa. This causes the diaphragm  24  to move in the second axial direction  46  forcing the spool  26  in the same direction. As a result, fluid from the discharge chamber  114  flows to the crankcase chamber  112  at an increased rate relieving the pressure in the discharge chamber  114 . The fluid in the discharge chamber  114  will drop to a pressure of 100 kPa. In each case, the equilibrium point of 50 kPa is maintained. 
     As presented above, providing a control valve with a diaphragm and associated elements described above presents numerous advantages. The diaphragm occupies significantly less volume than does the bellows. Therefore, the control valve may be manufactured significantly smaller as well. Also, bellows have a tendency to wear against opposing surfaces effecting the resiliency of a control valve. The diaphragm of the present invention, being smaller and constructed from a rigid material, does not wear against opposing surfaces. As a result, the diaphragm and control valve have a substantially longer useful life. 
     Although the present invention has been described and illustrated in detail, it is to be clearly understood that the same is by way of illustration and example only and is not to be taken by way of limitation, the scope of the present invention being limited only by the terms of the appended claims.