Abstract:
A method of controlling a pressure control valve assembly ( 69 ) to control a control pressure ( 43,71 ) in a system ( 11 ) including a source ( 47,49 ) of pressurized fluid. The control valve assembly is operable to vary the control pressure in response to variations in an electrical input signal ( 87 ) to an electromagnetic coil ( 81 ). The control valve assembly defines an inlet ( 77 ) in communication with the control pressure, an outlet ( 93 ) in communication with the low pressure, and a valve seat ( 89 ) disposed intermediate the inlet ( 77 ) and the outlet ( 93 ). A poppet valve member ( 91 ) is biased toward the valve seat by a biasing means including an armature ( 95 ), the biasing force being generally proportional to the input signal to the coil. The method comprises the steps of selecting a predetermined, minimum pressure for the control pressure corresponding to a minimum condition of the system. Next is selecting the biasing means to apply a biasing force to the poppet valve ( 91 ) corresponding to a fluid pressure less than the predetermined minimum pressure for the control pressure. The last step is selecting the electrical input signal ( 87 ) to correspond to the desired control pressure, an increasing input signal resulting in an increasing biasing force on the poppet valve member ( 91 ), but with no substantial movement thereof.

Description:
BACKGROUND OF THE DISCLOSURE  
       [0001]     The present invention relates to coupling devices of the type used to transmit torque, for example in a vehicle drive line, and more particularly, to a method of controlling the transmission of torque in and through such a coupling device.  
         [0002]     As used herein, the term “coupling device” will be understood to mean and include a device which is able to transmit torque from an input to one or more outputs, and in which there is a clutch assembly disposed between the input and the output, such that the amount of torque transmitted is a function of the extent of engagement of the clutch assembly. Within the scope of the present invention, the term “coupling device” means and includes both gear-type devices (such as differentials), as well as gearless-type couplings.  
         [0003]     Although the control method of the present invention may be utilized with many different types and configurations of coupling devices such as a coupling made in accordance with the teachings of U.S. Pat. No. 5,964,126 assigned to the assignee of the present invention and incorporated herein by reference, it is especially advantageous when utilized in conjunction with vehicle differentials of the type illustrated and described in U.S. Pat. Nos. 5,310,388 and 6,464,056, both of which are also assigned to the assignee of the present invention and incorporated herein by reference.  
         [0004]     In the differential coupling devices of the cited patents, there is a clutch pack operable to transmit torque between the input (housing connected to the ring gear) and the output (one of the axle shafts), with the degree of engagement of the clutch pack being determined by the fluid pressure in a piston chamber. The fluid pressure biases a clutch piston against the clutch pack. The differential coupling devices of the cited patents include a gerotor pump having one rotor fixed to rotate with the input and the other rotor fixed to rotate with the output, such that the flow of pressurized fluid into the clutch piston chamber is generally proportional to the speed difference between the input and the output. As used herein, the term “clutch pack” will be understood to mean and include both a multiple friction disc type clutch pack, as well as any of the other well known types of clutch assemblies, such as cone clutches, in which the degree of engagement is generally proportional to the fluid pressure acting on the clutch piston, or on an equivalent clutch-engagement structure.  
         [0005]     In differential coupling devices of the type described above, it is typical to provide a flow path from the clutch piston chamber to a reservoir or some other source of low pressure fluid, and to provide in this flow path some sort of control valve which can control the flow from the clutch piston chamber to the low pressure source, thereby controlling the pressure in the clutch piston chamber, and therefore, controlling the “bias torque”, i.e., the extent to which torque is transmitted from the input to the output.  
         [0006]     It has been known in prior art coupling devices of the type described above to provide a control valve of the type which comprises a flow metering type of device, in other words, a device which is essentially a variable flow restriction type of device. In such a device, fluid flows through a small orifice, with the pressure in the clutch piston chamber being determined by the “back pressure” created by the flow of fluid through the orifice. As has been recognized by those skilled in the art for some time now, one of the problems associated with this type of clutch pressure control is that the flow (and therefore, also the “clutch pressure”) is very much dependent upon factors such as fluid viscosity and fluid temperature which, in many devices of the type to which the present invention relates, will vary significantly during the period of operation of the device. Therefore, consistency of operation is difficult to achieve in coupling devices having the prior art type of pressure control system described above.  
         [0007]     Increasingly, vehicle differential coupling devices of the type to which the present invention relates are being utilized in conjunction with ABS (“Anti-skid Braking Systems”), and various other types of active ride control and vehicle dynamics control systems. When a vehicle differential coupling device is utilized in conjunction with such systems, it is critical to be able to vary the bias torque (i.e., the degree of torque transmission) at a rate of response which is not only very fast, but which is also predictable and repeatable.  
       BRIEF SUMMARY OF THE INVENTION  
       [0008]     Accordingly, it is an object of the present invention to provide an improved method of controlling the transmission of torque in a coupling device which overcomes the disadvantages of the above-described prior art.  
         [0009]     It is a more specific object of the present invention to provide a method of controlling the transmission of torque which achieves the above-stated object, but in which variations in the extent of torque transmission do not depend solely upon, nor are they caused by, variations in the rate of a fluid flow out of the clutch piston chamber.  
         [0010]     It is another object of the present invention to provide such an improved method of controlling torque transmission in which the clutch piston pressure is directly proportional to an electrical input signal, but in which no movement of a valve member is required in order to vary the clutch piston pressure, as the term “no movement” will be defined hereinafter.  
         [0011]     The above and other objects of the invention are accomplished by the provision of an improved method of controlling the transmission of torque in a coupling device including an input and at least one output, the coupling device comprising a housing defining a clutch cavity. A clutch assembly is disposed in the clutch cavity and includes a first clutch member fixed to rotate with the input, and a second clutch member fixed to rotate with the output. A clutch apply member is disposed in a pressure chamber and is operable to bias the first and second clutch members into torque-transmitting relationship in response to the presence of pressurized fluid in the pressure chamber. The coupling device includes a source of the pressurized fluid, and a control valve assembly operable to vary the fluid pressure in the pressure chamber, in response to variations in an electrical input signal transmitted to an electromagnetic coil. The control valve assembly defines an inlet in fluid communication with the pressure chamber, an outlet in fluid communication with a source of low pressure fluid, and a valve seat disposed intermediate the inlet and the outlet. A poppet valve member is biased toward the valve seat by a biasing means including an armature, the biasing force applied by the armature to the poppet valve member being generally proportional to the electrical input signal to the electromagnetic coil.  
         [0012]     The improved control method comprises the steps of selecting a predetermined minimum fluid pressure in the pressure chamber, corresponding to a predetermined minimum bias torque for the clutch assembly. The next step is selecting the biasing means to apply a biasing force to the poppet valve member corresponding to a fluid pressure less than the predetermined fluid pressure. The last step is selecting the electrical input signal to correspond to the desired fluid pressure in the pressure chamber and the desired bias torque for the clutch assembly, an increasing electrical input signal resulting in an increasing biasing force on the poppet valve member, but with no substantial movement thereof. 
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0013]      FIG. 1  is an axial cross-section of a vehicle differential coupling of the type with which the present invention may be utilized.  
         [0014]      FIG. 2  is an enlarged, fragmentary, axial cross-section, similar to  FIG. 1 , illustrating in greater detail that portion of the coupling device with which the method of the present invention is most closely associated.  
         [0015]      FIG. 3  is a perspective view of the valve housing portion of the coupling device shown in  FIGS. 1 and 2 , including an external, plan view of the control valve assembly utilized in the method of the present invention.  
         [0016]      FIG. 4  is an enlarged, axial cross-section of the control valve assembly shown in external plan view in  FIG. 3 . 
     
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0017]     Referring now to the drawings, which are not intended to limit the invention,  FIG. 1  illustrates a differential coupling, generally designated  11 , for use in a vehicle drive line, and which is of the general type illustrated and described in the above-incorporated patents, and which is especially well suited for use in connection with the control method of the present invention. The differential coupling device  11  comprises a housing, including a gear housing  13 , a clutch housing  15 , and a valve housing  17 , held together by any suitable means, well known in the art.  
         [0018]     Referring still to  FIG. 1 , the gear housing  13  defines a gear chamber  19 , and disposed therein, but by way of example only, there may be provided a typical differential gear set. In the subject embodiment, there is included a pair of input pinion gears  21 , rotatably mounted relative to a pinion shaft  23 , the pinion gears  21  being in toothed engagement with a pair of side gears  25  and  27 . In the subject embodiment, and again by way of example only, the input pinion gears  21  (along with the housings  13  and  15 ) may be considered the “input” to the coupling device  11 , while the side gears  25  and  27  comprises the “outputs” of the coupling device  11 . More specifically, for purposes of most of the subsequent description, the side gear  27  will be considered to comprise the “output” of the coupling  11 .  
         [0019]     The side gears  25  and  27  define sets of straight, internal splines  25 S and  27 S, respectively, which are adapted to receive right and left axle shafts (not shown herein), whereby the coupling device  11  transmits torque by means of the axle shafts to associated vehicle drive wheels (also not shown). It should be clearly understood, however, that the structure described hereinabove is by way of example only, and the control method of the present invention may also be used with various other types of structures. In fact, the control method of the present invention may be used generally to control a pressure control valve assembly, to control a control pressure, whether in a clutch piston chamber or not.  
         [0020]     Referring still primarily to  FIG. 1 , there is disposed within the clutch housing  15  a clutch pack, generally designated  29  which, as is shown in greater detail in  FIG. 2 , comprises a plurality of outer discs  31  which are in splined engagement with a set of internal splines defined by the clutch housing  15 . In addition, the clutch pack  29  includes a plurality of inner discs  33 , which are interleaved with the outer discs  31  in a well known manner, the inner discs  33  being in splined engagement with a coupling member  35 . The coupling member  35  defines a set of internal splines  35 S, which are also in splined engagement with the left axle shaft, such that the coupling member  35  is fixed, in the subject embodiment, to rotate with the side gear  27 .  
         [0021]     Also disposed within the clutch housing  15  is an annular housing insert  37  which cooperates with the adjacent coupling member  35 , and with the clutch pack  29 , to define a clutch cavity or clutch piston chamber  39 . Disposed within the clutch piston chamber  39 , and moveable axially therein, is a clutch piston  41  which cooperates with the housing insert  37  to define a piston pressure chamber  43 , which may better be seen in  FIG. 2 . As is well known to those skilled in the art of such devices, variations in the fluid pressure in the piston pressure chamber  43  will result in variations in the axial force applied by the clutch piston  41  to the clutch pack  29  and therefore, will result in variations in the “bias torque”, i.e., the torque transmitted through the clutch pack from the input of the coupling  11  to the output.  
         [0022]     Referring now primarily to  FIG. 2 , also disposed within the clutch housing  15 , and immediately to the left of the housing insert  37 , is a gerotor gear set comprising a stationary, eccentric member  45 , an internally-toothed outer rotor  47 , and an externally-toothed inner rotor  49 . The inner rotor  49  defines a set of straight, internal splines  49 S which are also in engagement with the left axle shaft, as described previously, such that the inner rotor  49  is fixed to rotate with the coupling member  35  and the side gear  27 . As is now well known to those skilled in the art of such devices, during normal, straight-ahead operation, the entire differential coupling  11  rotates as a unit, i.e., the housings  13  and  15  and the side gears  25  and  27  and the axle shafts all rotate at the same rotational speed. In that condition, there is no relative rotation between the outer rotor  47  and the inner rotor  49 , and therefore, there is no pumping of pressurized fluid from the volume chambers (formed between the teeth of the rotors  47  and  49  when they are relatively rotating).  
         [0023]     As is also well known to those skilled in the art, based in part upon the teachings of the above-incorporated patents, when there is differentiation, i.e., when there is a difference in the speed of rotation between the left and right axle shafts, there will also, of necessity, be a speed difference between the input (housings  13  and  15  and pinion gears  21 ) and the output (the left axle shaft). That speed difference between the input and the output will result in the rotation of the left axle shaft driving the inner rotor  49  which, in turn, will drive the outer rotor  47 , thus pumping pressurized fluid into an output chamber  51  from where it is communicated through an appropriate fluid port  52  in the housing insert  37 , such that the pressurized fluid is communicated into the piston pressure chamber  43 .  
         [0024]     Referring still primarily to  FIG. 2 , the housing insert  37  cooperates with the eccentric member  45  and the clutch housing  15  to define an axial fluid passage  53  which is in open communication with the piston pressure chamber  43 . The clutch housing  15  defines a radial fluid passage  55  which intersects the axial passage  53  and, at its radially inner extent, communicates with another axial fluid passage  57  which then, by means of a short radial passage  59 , communicates to the outer cylindrical surface formed by a hub portion  61  of the clutch housing  15 .  
         [0025]     Referring now primarily to  FIGS. 2 and 3 , it may be seen that the valve housing  17 , which is stationary within the outer differential housing (not shown herein), receives on its inner periphery, a pair of seal members  63  which are disposed on axially opposite sides of the radial passage  59 , the seal members  63  being in sealing engagement against the adjacent, outer cylindrical surface of the hub portion  61 . As may best be seen in  FIG. 3 , the valve housing  17  includes an inlet portion  65  which, preferably, extends down into a “source” of low pressure fluid, which would typically comprise a reservoir or sump containing fluid and disposed within the outer differential housing, as is well known in the art.  
         [0026]     The valve housing  17  also includes a port portion  67  on which is mounted a pressure control valve assembly, generally designated  69 , to be described in greater detail in connection with the description of  FIG. 4 . The valve housing  17  defines a generally radial passageway  71  (cut through at an angle to the plane of  FIG. 2 ) which extends radially inward to be in continuous fluid communication with the radial passage  59 , the radial passageway  71  having its radially outer end disposed within the port portion  67 . Therefore, the radial passageway  71  is disposed adjacent the pressure control valve assembly  69 , as is represented somewhat schematically in  FIG. 4 . Thus, it may be seen that, at any given point in time, the fluid pressure in the passageway  71 , which is effectively the “inlet” of the pressure control valve assembly  69 , is substantially identical to the fluid pressure in the piston pressure chamber  43 .  
         [0027]     Referring now primarily to  FIG. 4 , but in conjunction with  FIG. 3 , the pressure control valve assembly  69  will be described in some detail. The valve assembly  69  includes a valve body  73 , preferably comprising a ferromagnetic material, which comprises several different portions. The valve body  73  includes an inlet portion  75 , which is received within the port portion  67  of the valve housing  17 , and defines an inlet  77  in open fluid communication with the radial passageway  71 . Finally, the valve body  73  includes a reduced diameter portion  78 , which also serves as a “pole piece”, as that term is well understood by those skilled in the art of electromagnetic control valves.  
         [0028]     The larger diameter portion of the valve body  73  (i.e., the portion axially between the inlet portion  75  and the pole piece  78 ) supports a bobbin  79 , about which is wound an electromagnetic coil  81 , and the coil  81  is surrounded by, and encapsulated by, a generally cylindrical housing member  83 . Disposed within an open end of the housing member  83  is a cap member  85  which has formed integrally therewith, a wiring harness support  87 , shown only in  FIG. 3 . It will be understood by those skilled in the art that appropriate electrical input signals are transmitted to the electromagnetic coil  81  by means of wires which pass through the wiring harness support  87  and therefore, for simplicity of reference, the reference numeral “ 87 ” will also be used for the “electrical input signal” to the electromagnetic coil  81 . It should be understood by those skilled in the art that although reference will be made hereinafter to varying the “current” of the electrical input signal  87 , what would be more typical would be to vary the duty cycle of a pulse-width-modulated (PWM) type of signal, such that the RMS value of current would vary as the duty cycle is varied. However, the particular form of the input signal is not a limitation on the scope of the present invention.  
         [0029]     Referring still primarily to  FIG. 4 , disposed within the inlet portion  75  is a generally annular steel seat member  89 , against which is seated a poppet valve member comprising a steel ball  91 . Disposed adjacent the steel ball  91 , the valve body  73  defines one or more radial vent passages  93  which are in open fluid communication with a fluid reservoir, or fluid sump, defined by the outer differential housing (not shown herein). The fluid reservoir referred to hereinabove is the same reservoir or sump or “source” of low pressure fluid into which the inlet portion  65  extends, and draws fluid up into the valve housing  17 . Therefore, hereinafter, and in the appended claims, the inlet portion  65  will be referred to as the source of low pressure fluid, in view of the fact that the portion of the structure shown in the drawings which is most closely associated with the fluid reservoir is the inlet portion  65 .  
         [0030]     Disposed at the left end (in  FIG. 4 ) of the valve body  73  is a generally cylindrical armature member  95 , and seated between the cap member  85  and the armature member  95  is a compression spring  97 , biasing the armature member  95  to the right in  FIG. 4 . Received within an opening in the right end of the armature member  95  is a steel pin  99  which is fixed relative to the armature member  95 , and passes through an elastomeric member  101 . The elastomeric member  101  acts as both a fluid seal and a wiper, to wipe contamination particles off of the steel pin  99 , as the pin  99  occasionally reciprocates slightly relative to the elastomeric member  101 . The extreme right hand end of the steel pin  99  is disposed against the steel ball  91 , holding it against the seat member  89 .  
         [0031]     In the embodiment shown in  FIG. 4 , and described herein, the armature  95  and the compression spring  97  together comprise the “biasing means”, biasing the ball  91  against the seat member  89 . However, it should be understood that, within the scope of the present invention, the spring  97  may be left out of the pressure control valve assembly  69 , if the predetermined, minimum fluid pressure for the piston pressure chamber  43  is low enough that no spring bias is required. In that case, the armature  95  alone comprises the “biasing means” acting on the ball  91 , but if the desired minimum fluid pressure is high enough to require the compression spring  97 , then the “biasing means” comprises the armature  95  and the spring  97  acting together.  
         [0032]     In reference to  FIG. 4 , the seat member  89 , the ball  91  and the pin  99  have all been described as comprising “steel” members, although those skilled in the art of electromagnetic control valves will understand that, although steel is a preferred material, certain other materials may also be used. What is essential in regard to the items  89 ,  91 , and  99  is that the hardness of each part is selected, relative to the hardness of the other parts, so that neither the seat member  89 , nor the ball  91 , gets damaged as a result of the periodic dis-engagement and re-engagement therebetween. Another factor in the selection of the materials is the likelihood of corrosion being caused by the particular oil being used within the coupling. For example, brass is frequently used for parts such as the seat member  89  and the ball  91 , but the oil in the commercial, subject embodiment of the coupling  11  is corrosive to brass, and therefore, brass is unacceptable for these parts. It is believed to be well within the ability of those skilled in the art of hydraulic valves to select the appropriate materials for the items  89 ,  91  and  99 , to meet the needs of the particular application for the valve assembly.  
         [0033]     As was noted previously, the piston pressure chamber  43  is in relatively open, although tortuous, fluid communication with the radial passageway  71 , and therefore, with the inlet  77 . Therefore, whenever there is differentiation occurring within the coupling  11 , thus resulting in relative rotation between the outer rotor  47  and inner rotor  49 , pressurized fluid is pumped from the contracting fluid volume chambers through the output chamber  51 , through the fluid port  52 , and into the piston pressure chamber  43 . As differentiation continues, over even a very short period of time, the fluid pressure in the pressure chamber  43  will rapidly increase, subject to the operation of the pressure control valve assembly  69 , as will be explained subsequently. In the subject embodiment of the present invention, with the inner rotor  49  having six teeth (or lobes), fluid pressure in the pressure chamber  43  will increase substantially within one-sixth of a revolution of the rotor  49 . How long a period of time is required to build pressure will depend, in part, on the speed of differentiation between the input and the output, as is well known in the art of differentiating type devices.  
         [0034]     As is also well known to those skilled in the art, one of the key operating parameters of the coupling  11  is its “response time” which, in this context, means the time required for the coupling  11  to change (transition) from a disengaged condition (with the clutch pack  29  transmitting either substantially zero torque, or some predetermined minimum bias torque), to a particular, selected torque transmitting condition of the coupling  11  (in which a certain, predetermined bias torque is being transmitted by the clutch pack  29 ). In accordance with one aspect of the present invention, the fluid pressure in the pressure chamber  43  may be maintained at some predetermined minimum fluid pressure, corresponding to a predetermined minimum bias torque on the clutch assembly  29 . Typically, to achieve fastest response time, this predetermined minimum fluid pressure would be a pressure just below the pressure required to move the clutch piston  41  into effective engagement with the clutch pack  29 . Therefore, in the subject embodiment, when there is no electrical input signal  87  to the electromagnetic coil  81 , the total biasing force exerted against the poppet valve member (steel ball)  91  is determined solely by the biasing force of the compression spring  97 . However, by maintaining the fluid pressure in the piston pressure chamber  43  just below, or somewhat below, what is required to begin to engage the clutch pack  29 , the time for the coupling to “respond” (i.e., to transition the clutch pack  29  from disengaged to engaged), is greatly reduced., and the overall response of the system is substantially improved.  
         [0035]     In the “zero torque bias” condition being selected (by virtue of the electrical input signal  87  being “off”), if there is sufficient differentiation to rotate the rotors  47  and  49  and pump sufficient fluid into the pressure chamber  43 , so that the pressure chamber  43  would rise above the predetermined minimum fluid pressure (needed to engage the clutch pack  29 ), that same fluid pressure, also present in the inlet  77 , will overcome the force of the compression spring  97  and “lift” the poppet valve member  91  (move it slightly to the left in  FIG. 4 ) just enough to bleed or vent the pressurized fluid past the seat member  89 , and out through the radial passages  93  to the reservoir or sump as described previously. The poppet valve member  91  will re-seat against the seat member  89  when an equilibrium condition is re-established between the predetermined minimum fluid pressure in the chamber  43  (and in the inlet  77 ) and the force of the compression spring  97 . In other words, “equilibrium” is achieved whenever the force exerted on the ball  91  by the spring  97  is equal to the force exerted on the ball  91  by the fluid pressure in the inlet  77  acting on the ball  91 , i.e., over the area within the inside diameter of the seat member  89 .  
         [0036]     When it is desired to increase the torque bias of the coupling  11 , the electrical input signal  87  is merely increased to some predetermined current level (corresponding to the desired torque bias), such that the total force on the poppet valve member  91  becomes the sum of the force of the compression spring  97 , and the force exerted by the armature member  95  and steel pin  99  (i.e., under the influence of the electromagnetic force from the coil  81 ). In accordance with an important aspect of the present invention, it is thus possible progressively to increase the fluid pressure in the piston pressure chamber  43  simply by increasing the electrical input signal  87 , but this increased “setting” occurs without any substantial movement of the poppet valve member  91 , and requires no flow through the valve (or change in a rate of flow, etc.). Instead, the pressure in the chamber  43  increases in response to an increase in the force tending to seat the poppet valve member  91 , as it remains seated. Thus, it has been stated herein, and in the appended claims, that the pressure in the pressure chamber  43  is increased, but with “no substantial movement” of the poppet valve member  91 .  
         [0037]     As a result, the performance and response time of the coupling  11 , and the pressure control valve assembly  69 , are relatively independent of factors such as the fluid viscosity and fluid temperature, factors which typically cause substantial variations in the performance and response time of the prior art pressure control valves and clutch control systems.  
         [0038]     What has been stated above in regard to increasing the fluid pressure in the chamber  43  by increasing the current to the coil  81  also applies generally in the reverse condition, i.e., as the current to the coil  81  is decreased from a relatively high current level, the total force exerted on the poppet valve member  91  also decreases, and therefore, the pressure in the piston pressure chamber  43  decreases. However, as a practical matter, the decreasing pressure in response to a decreasing current to the coil  81  does require just enough movement of the poppet valve member  91  from the seat to bleed enough fluid from the inlet  77  to the reservoir to reach the reduced pressure level desired. As will be understood by those skilled in the art, from a reading and understanding of the foregoing specification, if, at any time, the pressure in the inlet  77  exerts a force against the steel ball  91  (i.e., against the area of the ball  91  within the ID of the seat member  89 ), which is greater than the total force of the biasing means, i.e., force exerted on the ball  91  by the spring  97  and the electromagnetic coil  81 , the ball  91  will lift slightly from its seat  89  and relieve just enough fluid pressure (through the vent passages  93 ) to re-establish the equilibrium condition between the opposing forces acting on the ball  91 . Thereafter, the ball  91  will again be seated, and remain seated as long as the equilibrium condition remains.  
         [0039]     Referring again primarily to  FIG. 4 , it may be seen that between the left end of the reduced diameter portion  78  and the right end of the armature member  95 , there exists an air gap  103 , as is generally well known to those skilled in the solenoid valve art as the “working air gap”. However, in accordance with an important aspect of the invention, the air gap  103  remains substantially “constant” (i.e., the axial “gap length” in  FIG. 4 ) because of the fact that pressure changes (in the pressure chamber  43 ) occur without substantial movement of the poppet valve member  91  (and therefore, without substantial movement of the armature member  95 ). The constant air gap  103  (even as the electrical input signal  87  and the pressure in inlet  77  vary) makes it much more feasible to provide a predetermined, known relationship of “control” pressure (in inlet  77  and pressure chamber  43 ), as a function of electrical current (to the coil  81 ). The constant working air gap  103  determines the magnetic calibration of the pressure control valve assembly  69 , because the air gap  103  determines, for a given current applied to the coil  81 , the electromagnetic force applied to the armature, and thus, the force applied to the pin  99  and the ball  91 . The pin  99  is pressed into the armature  95  to a specific, predetermined depth as part of the process of setting the calibration, as will be explained further subsequently.  
         [0040]     During the assembly process, a measurement is taken from the “top” of the steel ball  91  (i.e., the leftmost end in  FIG. 4 ) to the “top” of the pole piece  78  (i.e., the leftmost end in  FIG. 4 ). To that measured distance is added the desired amount of the working air gap  103 , and then the steel pin  99  is pressed into the armature  95  until the pin extends from the armature a distance equal to the sum of the measured distance plus the desired working air gap  103 . After the above calibration-and-assembly step has been performed, the working air gap  103  should be substantially equal to the desired gap. By way of example only, in the intended, commercial version of the pressure control valve assembly  69 , the working air gap  103  is about 0.009 inches (about 0.228 mm).  
         [0041]     If the desired working air gap  103  is achieved, then it is possible also to achieve the predetermined, known relationship of control pressure in the piston pressure chamber  43  (and inlet  77 ) to the input signal  87  to the electromagnetic coil  81 , without any further calibration or adjustments.  
         [0042]     One further step in the calibration-and-assembly of the pressure control valve assembly  69  is rotating the cap member, in the absence of a current to the coil  81 , until the force of the spring  97  on the ball  91  is in equilibrium with a known, predetermined fluid pressure in the chamber  43  and the inlet  77 . As explained previously, this known, predetermined fluid pressure would, in a coupling of the type shown herein (but by way of example only) correspond to a “standby” pressure to be maintained in the piston pressure chamber  43  when it is intended for the clutch pack  29  to be disengaged, but with the fluid pressure in the chamber  43  being maintained at a pressure high enough that engagement of the clutch pack  29  can be accomplished very quickly. Of course, in a given application for the pressure control valve assembly  69 , if the standby pressure is to be at nearly zero pressure, the compression spring  97  would be left out of the assembly  69 , as discussed previously, in which case the calibration step described above would be eliminated from the calibration-and-assembly process.  
         [0043]     The invention has been described in great detail in the foregoing specification, and it is believed that various alterations and modifications of the invention will become apparent to those skilled in the art from a reading and understanding of the specification. It is intended that all such alterations and modifications are included in the invention, insofar as they come within the scope of the appended claims.