Fluid flow regulator

A fluid flow regulator including a servo valve assembly and a throttle valve assembly. The servo valve assembly includes a pair of double acting actuators, biased in one direction by an adjustable spring, contoured ports and piping passageways to the ports and to other chambers of the assembly. The actuators are effectively connected together and function to oppose each other. The chambers formed between the two actuators and the respective cylinder chambers formed on the exterior ends of each actuator are used for pressure sensing purposes. Each individual actuator may be comprised of up to three piston lands with adequate separation between to form chambers for fluid transfer purposes. Any leakage or servo flow entering these chambers would be communicated by passageways to appropriate interface ports on the regulator. Due to its sensitivity, the servo valve assembly effects a desired action by variably restricting the flow or pressure in a passageway controlled by the throttle valve assembly. This result is achieved by throttle valve control.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
The present invention relates to a device for regulating fluid pressure 
and/or flow. These devices are generally referred to as fluid flow 
regulators. The pressure regulators will regulate flow to maintain a 
desired set pressure level or magnitude. The flow regulators, in turn, 
regulate pressure to maintain set fluid flowrates. 
2. Prior Art 
Heretofore, regulators used thin membrane devices such as diaphragms or 
bellows for pressure level sensing purposes to initiate required 
operational action. Because of their structural strength, these membrane 
devices effectively limit such regulators to process applications which 
operate at low or moderate pressure levels. Additionally, these membranes 
are usually subjected to relatively high fatigue stresses because of their 
flexing mode of operation and are generally the chief cause of regulator 
failure. 
An example of such a regulator is disclosed in my U.S. Pat. No. 3,143,134. 
OBJECTS AND SUMMARY OF THE INVENTION 
It is an object of the present invention to provide a fluid regulator which 
can reliably provide accurate, sensitive and responsive regulation and can 
be used in exceptionally high pressure process applications; 
It is a related object of the present invention to provide a fluid 
regulator which is readily adaptable for use as a flow regulator and a 
back pressure regulator. 
The present invention essentially comprises two major sub-assemblies and a 
combination of such sub-assemblies. The sub-assemblies comprise, 
respectively, a first means sensitive to and operative by fluid pressure 
conditions to effect restricting the servo flow from a first passageway 
connected to a second means and thereby operating and controlling the 
second means to effect a desired action by variably restricting the flow 
or pressure in a second passageway. The first means hereafter is referred 
to as the servo valve assembly and the second means is referred to as the 
throttle valve assembly. 
The servo valve assembly is basically comprised of a pair of double acting 
actuators, a spring, a spring adjusting means, contoured ports and piping 
passageways to the noted ports and to other chambers of the assembly. The 
actuators, each with slightly different effective areas, are housed in 
close fitted cylinder bores and are effectively connected together and 
function to oppose each other. The chambers formed between the two 
actuators and the respective cylinder chambers formed on the exterior ends 
of each actuator are used for pressure sensing purposes. Each individual 
actuator may be comprised of up to three piston lands with adequate 
separation between to form chambers for fluid transfer purposes. Any 
leakage or servo flow entering base chambers would be communicated by 
passageways, interconnected or otherwise, to appropriate interface ports 
on the regulator. Subsequent connection to these interface ports would 
depend upon the application and the operating fluid used: 
for applications using air or other environmentally compatible fluids 
capable of being expended into the atmosphere, individual passageways, 
each terminating at appropriate exhaust ports, could be used to drain and 
dispose the leakage and servo flow from the respective chambers. 
For liquid applications, interconnecting passageways terminating at a 
single interface port could be used. External piping would be necessary to 
transport the leakage and servo flows to a vented tank or to some other 
low pressure source.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
In FIG. 1, servo valve assembly 10 is comprised of servo valve 20, spring 
12, spring adjustment means 60, servo port 14, relief port 16, and piping 
passageways provided or installed in regulator body 2 as shown. Servo 
valve 20 is comprised of actuator 30 connected by rod 35 to actuator 40. 
Actuator 30 is itself comprised of piston land 32 connected by rod 39 to 
piston land 34. Actuator 40 is comprised of piston lands 42, 44 and 46 
connected by rods 49 and 47 respectively, as shown. The chambers formed 
between the respective piston lands of the two actuators are 
interconnected by passageways 52, 54 and 56 and terminate at an interface 
leakage port 58. The external chambers of the two actuators are 
interconnected by passageways 55 and 57 and terminate at the interface 
reference pressure port 59. The chamber formed between actuators 30 and 40 
is interconnected with relief port 16 and passageway 86 of the throttle 
valve assembly 100 by passageways 51, 53 and 88. The servo port 14 is 
interconnected with chamber 77 of the throttle valve assembly 100 by 
passageway 89 as shown. 
Consider that a pressure source P2 is applied to the chamber between the 
two actuators 30 and 40 and that it is at a higher level than the pressure 
source Pr connected to interface leakage port 59, which is applied into 
the exterior chambers of both actuators. Also consider that pressure 
source Pa is connected to leakage port 58 and is applied into the 
connected chambers formed between the respective piston lands of each 
actuator. With servo valve 20 arranged as shown in FIG. 1, i.e., with 
spring 12, compressed at a fixed deflection setting, applying a 
mechanically developed force on piston 42 of the larger actuator 40, and 
with pressure sources applied to the respective chambers noted above, an 
equation describing the forces acting on servo valve 20 would be as 
follows: 
EQU PrA1-PaA2+PaA2-P2A3+P2A4-PaA5+PaA5-PaA6+PaA6-PrA7-FS=0 
WHERE: 
Pr=Fluid pressure applied in the external chambers of actuators 30 and 40 
Pa=Fluid pressure applied in the chambers between the pistons of respective 
actuators 30 and 40 
P2=Fluid pressure applied in the chamber between actuators 30 and 40 
A1=Effective area of piston 32 exposed in the external chamber of actuator 
30 
A2=Effective areas of pistons 32 and 34 exposed in the chamber between said 
pistons 
A3=Effective area of piston 34 exposed in the chamber between actuators 30 
and 40 
A4=Effective area of piston 46 exposed in the chamber between actuators 30 
and 40 
A5=Effective areas of pistons 46 and 44 exposed in the chamber between the 
said pistons 
A6=Effective areas of pistons 44 and 42 exposed in the chamber between the 
said pistons 
A7=Effective area of piston 42 exposed in the external chamber of actuator 
40 
FS=Spring force applied by spring 12 
Combining and rearranging, the equation reduces to: 
EQU FS=P2(A4-A3)-Pr(A7-A1) 
considering that the respective piston areas are mechanically constant, the 
equation can be characterized as: 
EQU FS=aP2-bPr 
It thus can be seen that servo valve stability is achieved by the proper 
relationship of the spring force and the two pressure sources Pr and P2. 
If the pressure regulator is referenced to atmosphere, i.e., Pr is 
atmospheric pressure, the equation can be further characterized as: 
EQU FS=aP2-k 
Any change in the two variables, FS and P2, would unbalance the servo 
valve's equilibrium and would cause it to move. The amount of movement due 
to a change in pressure P2 may be determined from: 
##EQU1## 
where: .DELTA.P2=the change in pressure level P2 from one instant to 
another. 
This indicates that the sensitivity of the servo valve 20 is dependent only 
upon the characteristics of spring 12. In comparison, regulators using 
diaphragms, bellows or similar functioning items must additionally include 
the spring rates of such components to that of the spring in any movement 
equation pertaining to the particular regulators pressure sensing 
capabilities. 
The opening area of servo port 14, contoured or otherwise, is restrictable 
by metering edge 44a of piston land 44. Thus any movement of servo valve 
20 opens or closes servo port 14 and thereby can vary or control fluid 
flow from passageway 89. 
It is thus shown that the servo valve assembly 10 is sensitive to the 
pressure level in passageway 86 and reacts to any variation in that level 
to vary the fluid flow from passageway 89. 
The incorporation of relief port 16 and related passageway 51 provides a 
beneficial means of venting detrimental leakage flow from the communicated 
passageway 86 under certain operating conditions and will be described 
further hereinafter. In addition, a physical and functional relationship 
exists between relief port 16 and servo port 14. When metering edge 44a of 
actuator 40 is in line with and effectively covers or seals servo port 14, 
metering edge 34a of piston land 34 slightly covers, or, as a minimum is 
in line with the edge of, relief port 16. Functionally, relief port 16 is 
covered by piston land 34 before the servo port 14 is opened by piston 
land 44 when the servo valve is moving in that particular direction. 
Conversely, servo port 14 is covered before relief port 16 is opened when 
moving in the opposite direction. 
The configuration of servo valve assembly 10 has a certain symmetry which 
enables it to be readily rearranged to create a new configuration which 
reverses the servo valves direction of motion for a like pressure change. 
As shown in FIG. 1, the application of pressure source Pr to the external 
chambers of actuators 30 and 40 and the application of the higher level 
pressure P2 to the chamber between the two actuators results in a 
developed force which would move the servo valve 20 to the right toward 
the larger actuator. If the application of the two pressure sources are 
reversed, i.e., P2 applied in the external actuator chambers and Pr in the 
chamber between the two actuators, the resultant force would move the 
servo valve to the left toward the smaller actuator. It should be apparent 
that an alternate servo valve could thus be created by relocating the 
biasing spring and related adjustment means to act on the smaller actuator 
and by relocating the servo and relief ports as necessary to function as 
required. The servo valve assembly shown in FIG. 2 uses this arrangement. 
The utilization of the chambers between the respective pistons in each 
actuator provide a beneficial feature. These chambers intercept and 
exhaust any leakage passing the close fitted piston lands and prevent 
intermixing of the critical pressure sources P2, Pr and the pressure 
source in passageway 89, which will be subsequently identified as P3. This 
interception eliminates any inaccuracy in sensed pressure level that could 
result from such intermixing of pressure sources in any critical chamber 
having one of these applied sources. 
Many applications do not require regulators with the degree of accuracy 
possible with the design shown in FIG. 1. FIG. 2 shows schematically a 
more economic servo valve assembly configuration which, while not 
attaining in absolute terms the accuracy potential of the servo valve 
shown in FIG. 1, will provide similar objectives and regulating features. 
With pressure sources applied as shown in FIG. 2, leakage will flow into 
chamber 146 from the exterior chambers 142 and 144 by passing through 
diametral clearances between each of the two piston lands 122 and 124 and 
their respective cylinder bores 132 and 134. Servo flow (to be 
subsequently identified) from the throttle valve assembly 200 enters 
chamber 146 through servo port 114. The volume of chamber 146 and the 
dimensions of passageway 148 and exhaust port 149 are sized and configured 
to drain and exhaust the combined leakage and servo flow without causing 
any significant pressure level rise in chamber 146 from that experienced 
in the external pressure source Pr. 
It should be noted that, because of the labyrinth nature of close fitted 
pistons and bores, the leakage flowrate would be small and relatively 
constant over a wide range of levels between pressures P1 and Pr. Also, 
the nominal rate of servo flow during operation can be made relatively 
small since it is dependent upon sizing the orifice 162 in passageway 164 
to attain a desired response of the throttle valve assembly which will be 
subsequently discussed. 
THROTTLE VALVE ASSEMBLY 
In FIG. 1, throttle valve assembly 100 is comprised of piston 70 installed 
in suitable closely fitted bores 76 and 78 of regulator body 2, a throttle 
port 81, an orifice 73 and related passageways. Piston 70 has two unequal 
diameter piston lands: a motor piston land 72 and a throttle piston land 
74. The two piston lands are connected by a suitable rod 71. Throttle port 
81 at the interior end of regulator inlet passageway 82 is opened and 
closed by the throttle piston land 74. Thus when fluid is supplied to the 
inlet passageway 82 at sufficient pressure and flow capacity, a flowrate 
is developed through any existing opening of the throttle port 81 and is 
controllable by any subsequent repositioning of throttle piston land 74. 
This flow then passes through valve chamber 79 to the exit passageway 86. 
Passageway 84 communicates inlet passageway 81 to valve chamber 75 formed 
between piston lands 72 and 74. Orifice 73 is shown forming an opening in 
the wall of piston land 74 and thereby communicates chamber 75 with servo 
pressure chamber 77. This orifice could alternatively be placed in the 
piping passageway connecting the two chambers as illustrated in the back 
pressure regulator depicted in FIG. 2. This orifice allows fluid from the 
high inlet pressure source in chamber 75 to flow into chamber 77. Chamber 
77, in turn is communicated with servo port 14 of servo valve assembly 10 
by passageway 89 as previously noted. 
It is apparent that if fluid flow, hereafter called servo flow, from 
chamber 77 is restricted in its escape through passageway 89, the pressure 
level in chamber 77 will approach that in chamber 75. The reverse is also 
true, i.e., if the flow is unrestricted through passageway 89, the 
pressure level in chamber 77 will be much lower than the inlet pressure in 
chamber 75. Orifice 73 is sized to attain a desired regulator response 
characteristic which is also dependent upon the characteristics of the 
fluid used, the mass of piston 70, the volume of valve chamber 77, the 
masses of servo valve 20 and spring 12 and the opening area 
characteristics of servo port 14. 
For the purposes of the following discussion, the pressures existing in and 
around the throttle valve assembly are as follows: 
P1=Inlet pressure of fluid entering passageway 82 
P2=Pressure leaving through passageway 86 
P3=Pressure of fluid in servo pressure chamber 77 
Critical areas of piston 70 shown in FIG. 1 are as follows: 
A1=Effective area of throttle piston land 74 
A2=Effective area of motor piston land 72 
A3=Effective area of connecting rod 71 
The forces acting on the throttle piston 70 are as follows: 
EQU P2A1-P1(A1-A3)+P1(A2-A3)-P3A2=0 
or 
EQU P3A2=P2A1+P1(A2-A1) 
Since areas A1 and A2 are constant because of their mechanical 
construction, it can be seen that piston 70 is moved or is stationary 
depending upon the relationship of the levels of the three pressures P1, 
P2 and P3. Any change in any of the pressures would unbalance the 
equalization forces acting on piston 70 and thus cause it to move in the 
direction of the resultant unbalanced force until one or both remaining 
two pressures are changed to again provide an equalization force on piston 
70 such that the above equation is satisfied. 
PRESSURE REGULATOR OPERATION 
When passageway 88, interconnected with relief port 16 and the applicable 
servo valve chamber, is terminated and communicates with exit passageway 
86, the pressure regulator illustrated in FIG. 1 will vary and control 
flow as necessary to maintain the flowrate exiting passageway 86 at a set 
or regulated pressure level as established by the force setting applied to 
actuator 40 by spring 12. The regulator may be considered to function in 
the following manner. Consider the following conditions exist: 
1. Reference pressure port 59 is suitably connected to a vented tank and 
communicates atmospheric pressure to the respective exterior chambers of 
actuators 30 and 40. 
2. Leakage port 58 is connected by separate piping means to a vented tank 
or to another system pressure source such that a pressure level less than 
or equal to that of the atmosphere is developed in the respective chambers 
between the piston lands of actuators 30 and 40 under all operating 
conditions. 
3. The residual compression of spring 12 positions servo valve 20 to be in 
physical contact with its travel stop 13 such that servo port 14 is fully 
opened and relief port 16 is fully closed. 
Consider initial operation when fluid with sufficient flow capacity is 
suitably applied to inlet passageway 82. The initial surge of fluid 
passing through throttle port 81 causes the level of pressure P1 to 
increase in passageway 82 and in chamber 75. Two simultaneous actions 
result: First, piston 70 moves to fully open throttle port 81 and second, 
a servo flowrate is initiated through orifice 73 passing through the fully 
opened servo port 14 and ultimately to the external vented tank through 
leakage port 58. Full opening of the throttle port 81 accelerates the 
flowrate passing through the regulator exit passageway 86 and into the 
suitable connected downstream system. When flow reaches a restriction 
which causes the pressure P2 to rise, the increase is reflected in 
passageway 86 and in the communicated chamber between actuators 30 and 40. 
This causes servo valve 20 to move to the right and to start closing servo 
port 14, which in turn reduces servo flow and increases the level of 
pressure P3 in chamber 77. This causes piston 70 to move and to start 
closing throttle port 81 reducing flow through the regulator. A state of 
equilibrium could be said to exist when the level of pressure P2 ceases to 
increase and becomes stable and when both, servo valve 20 and throttle 
piston 70 cease in their movements. At this time, the force developed by 
regulated outlet pressure level P2 satisfies the spring force setting 
acting on servo valve 20 per the equation: 
EQU FS=aP2+k 
The servo valve in turn restricts servo flow to develop the required P3 
pressure level which satisfies the throttle valve force equation: 
EQU P3A2=P2A1+P1(A2-A1) 
The position of metering edge 44a of piston land 44 relative to the servo 
port 14 at this time is known as the null position of the servo valve. The 
position of throttle piston land 74 relative to throttle port 81 
establishes the required port area opening which, by being subjected by 
the difference in level in pressures P1 and P2, develops the flowrate 
which attains the regulated pressure level P2 in the external system 
restriction. 
Now consider an operational requirement whereby the flowrate is to be 
increased by increasing pressure level P2 by regulator adjustment. Using 
the spring adjustment means 60, spring 12 is further compressed increasing 
the force applied to servo valve 20 and causing it to move from the null 
position and further open servo port 14. This increases the servo flow 
reducing pressure level P3 in chamber 77 and unbalancing piston 70. The 
piston moves to open the throttle port 81 causing an increased flowrate 
through the regulator. The regulator regains its equilibrium state at a 
new throttle piston setting when the regulated pressure P2 reaches the 
newly set level which causes servo valve 20 to return to its so-called 
null position with the higher spring force applied. 
Now consider the condition whereby the restriction in the external system 
is increased causing pressure level P2 to increase. As pressure level P2 
increases, two simultaneous regulator actions occur which integrate and 
accelerate the response rate of the regulator. First, servo valve 20 is 
unbalanced causing it to move from the null position to close the servo 
port 14. Second, throttle piston 70 is similarly unbalanced causing it to 
move to open throttle port 81 and to increase the flowrate through the 
regulator. This evidently acclerates the pressure rise sensed by servo 
valve 20 causing its faster movement to close servo port 14. Closing of 
the servo port 14, as previously described, reduces flow through the 
regulator until a new equilibrium state is again reached and servo valve 
20 returns to a null position and the set pressure level P2 is regained. 
Now consider the condition whereby all flow is stopped in the downstream 
system. Regulator operation is as discussed previously, except piston 70 
continues movement until it closes throttle port 81. The pressure level P2 
at that instant would be within the sensitivity band of the spring 
adjusted pressure level setting. Leakage, however minute, initiates from 
throttle port 81 passing through diametral clearances between piston 72 
and cylinder bore 78 enters exit passageway 86 and increases pressure 
level P2. This causes servo valve 20 to continue moving and eventually 
opens relief port 16 which was covered by piston land 34. Movement 
continues until an equilibrium condition develops whereby a relief port 
opening is reached which satisfies the leakage rate entering exit 
passageway 86. The level increase in pressure P2 at this condition is a 
function of the added spring deflection from the servo valve null position 
to the one that satisfies the relief port opening setting. 
Any subsequent adjustment reducing spring force will cause a further relief 
port opening, venting the pressurized P2 passageways and ultimately 
stabilizing at a lower P2 level when equilibrium is reestablished. This 
relief port feature is also beneficial in that it will vent any pressure 
surge whenever fluid flow is drastically stopped in an external system. 
This would possibly eliminate hammer effects in the system. 
SPRING ADJUSTMENT MEANS 
The regulator configuration described above could readily be adapted to 
utilize any number of spring adjustment means. As shown in FIG. 1, spring 
adjustment means 60 is a motor operated gear driven screw arrangement 
which traverses a non-rotating spring seat along a guide in the cylinder 
to change the length of spring 12. Objectively, in the broadest sense, the 
regulator designs described herein may use any spring adjustment means 
that can be activated and/or operated manually, electrically, optically, 
pneumatically, hydraulically or by any other means, to compress, position 
and/or to deflect the biasing spring in the servo valve assembly. 
BACK PRESSURE REGULATOR 
In FIG. 2 is described a regulator for controlling pressure level upstream 
rather than downstream of the regulator in contradistinction to the 
regulator in FIG. 1. This embodiment of the invention also differs from 
that in FIG. 1 in that: 
1. The orifice through the wall of the motor piston land is replaced with 
an orifice 162 in passageway 164 interconnecting chamber 151 with servo 
pressure chamber 153 in throttle valve assembly 200. 
2. Throttle port 181 instead of communicating with the regulator inlet 
passageway is connected to the exit passageway 186. This necessitates 
providing passageway 176 to transmit pressure P2 to chamber 179 and to act 
on the appropriate area of throttle piston land 174. 
3. The servo valve assembly is changed as follows: 
a. the spring adjustment means is changed from a motor operated type to the 
manually operated screw type adjustment means 160, 
b. the servo valve assembly is changed to eliminate the leakage chambers 
and related passageways between piston lands on each actuator as 
previously described in the servo valve discussion, 
c. the passageways transmitting pressure sources P2 and Pr are interchanged 
and spring 112 is relocated to apply force to the smaller actuator instead 
of the larger as previously described. 
System flow which is to be regulated at a set pressure level is suitably 
applied to inlet passageway 182 and passes through chamber 151a and the 
resultant port 181 opening "throttled" by piston 174 and thence through 
passageway 186 to a suitably connected vented tank or other low pressure 
source. Passageway 188 transmits regulated pressure P1 to the exterior 
chambers 142 and 144 of servo valve assembly 110. 
Passageway 164 with incorporated orifice 162 communicates inlet passageway 
182 with throttle valve chamber 153 and with servo port 114. Servo flow 
developed through any opening of servo port 114 passes through chamber 146 
and passageway 148 and is exhausted to atmosphere through interface port 
149. 
It will be observed from the above description of the back pressure 
regulator shown in FIG. 2 that the regulator will serve to maintain the 
pressure level of the system fluid source applied to passageway 182 as a 
function of the compression to which spring 112 is adjusted. A pressure 
level greater than the set level would cause the servo valve to move 
against the set spring force closing servo port 114 and causing throttle 
valve 150 to move and open throttle port 181 and venting the pressure in 
passageway 182 until equilibrium is regained. Reverse operation would be 
experienced if a lower pressure level than set level was sensed in 
passageway 182. 
It will also be observed that the regulator defined in FIG. 1 could readily 
be converted into a back pressure regulator by simply: 
1. Relocating passageway 88 and connected passageway 53 to communicate with 
inlet passageway 82, 
2. Relocating servo port 14 to be throttled by piston metering edge 44b 
instead of 44a, and 
3. Eliminating relief port 16 and connected passageway way 51. 
FLOW REGULATOR 
The general equation for volumetric flowrate passing through an orifice is: 
EQU Q=kA.sqroot..DELTA.P 
where: 
Q=Flowrate, volume per time 
k=a constant 
.DELTA.P=Pressure loss across orifice, force per area 
Volumetric flow is therefore dependent upon the two variables, orifice area 
and pressure levels across the orifice. It is also evident that a constant 
flowrate would result if a constant pressure loss is maintained across a 
fixed orifice area. 
A flow regulator embodying the invention is shown in FIG. 3. This regulator 
will maintain or schedule a desired flowrate at the discharge side of the 
regulator in relation to an orifice area setting, fixed or varying, and in 
relation to the maintenance of a pressure loss setting, fixed or varying, 
across the orifice. 
Flow enters in the direction of the arrow in passageway 282 and leaves the 
regulator through exit passageway 286. This flow regulator uses the same 
components previously described above in connection with the pressure 
regulators depicted in FIG. 1 and FIG. 2, and for this reason, a 
structural description thereof would be repetitious and unnecessary. The 
flow regulator includes, however, several additional features not present 
in the previously described pressure regulators, namely, an orifice 292 in 
passageway 286 downstream of interconnecting P2 pressure sensing 
passageway 288, an orifice area adjustment means 290, and passageway 289 
which communicates passageway 286 downstream of orifice 292 with the 
exterior "reference" pressure chambers 242 and 244 of servo valve 220. 
With this arrangement of orifice 292 and passageway 289, the servo valve 
220 will sense the pressure loss or drop caused by the flow through 
orifice 292 and will endeavor to maintain a definite pressure loss, as set 
by the compression of spring 212, by the throttling action of throttle 
piston 270. Pressure level variations upstream or downstream of the 
regulator will not affect the delivery of the scheduled flowrate provided 
the difference between the inlet source pressure level in passageway 282 
and the exit pressure level leaving the regulator is, at any instant, 
always greater than the pressure loss setting across the orfice. 
Orifice area adjustment means 290 depicted in FIG. 3 is simply a manually 
operated screw driven tapered shaft which can traverse within the center 
of the orifice thereby varying the area setting by changing the relating 
distance between the tapered surface and the orifice parimeter. As in the 
case for the servo spring adjustment means, the flow regulator designs 
described herein can be adapted to use any number of orifice area 
adjusting techniques. Any means, operated or powered manually, 
electrically, hydraulically and et cetera, can be used provided that the 
resultant area setting is structurally stable and physically unaffected by 
the flow or pressures imposed. 
The flow regulator illustrated in FIG. 4, in contradistinction to that in 
FIG. 3, regulates fluid flow rate at the upstream side of the regulator 
and is more suitable for applications using incompressible fluids 
pressurized by positive displacement pumping elements. This regulator uses 
alternate throttle valve and servo valve design configurations for those 
shown in FIG. 3 and these have been previously described in the FIG. 1 and 
FIG. 2 pressure regulator discussions. The regulator in FIG. 4 differs 
from that in FIG. 3 mainly by the addition of a regulator inlet passageway 
360, a regulator exit passageway 362, by the inclusion of fixed area 
orifice 392 between the two passageways, and relocation of other pertinent 
passageway as noted herein. A branch passageway 382 communicates inlet 
passageway 360 to chamber 371 formed between respective piston lands of 
throttle piston 370. Passageway 388 communicates passageway 360 to servo 
valve chamber 346 between the two respective actuators 330 and 340. 
Passageway 357 communicates passageway 362 to the exterior servo valve 
chambers 342 and 344. Passageway 384 communicates the three leakage 
chambers between the respective piston lands on actuators 330 and 340 to 
the return passageway 386. 
The regulator in FIG. 4 functions generally in the following manner. 
Unregulated flow, at a rate capacity always greater than that required by 
the downstream using system, is suitably applied into inlet passageway 
360. Regulated flow at a set rate passes through orifice 392 and is 
discharged through exit passageway 362 to a suitably connected system 
utilizing such flowrate. Excess flowrate in passageway 360 is transmitted 
by branch passageway 382 through the throttle valve and into return 
passageway 386 and thence to a suitably connected vented tank or other low 
pressure source. 
It will be observed from the above description of the flow regulator in 
FIG. 4, that the regulator will serve to maintain a set pressure loss 
across the fixed area orifice as a function of the compression of spring 
312 and thus provide a set flowrate through the exiting passageway 362. 
It should also be observed that the single orifice shown in FIG. 3 and FIG. 
4 can be replaced with two or more independently operated adjustable type 
orifices arranged in series and/or parallel but within the confines 
between the two sensing passageways communicated to the respective servo 
valve chambers. Such regulator would provide or schedule a flowrate based 
on the effective area of the total number of orifices in the system 
without being affected by pressure variations or fluctuations upstream or 
downstream of the regulator. 
A possible application for such a regulator could be a fuel control for a 
jet engine. An adaptation of the regulators used in FIG. 3 or FIG. 4 could 
be devised using at least three area controlled orifices in series in the 
main passageway between the two interconnecting passageways sensing the 
pressure levels across the three orifices. One primary orifice area is 
scheduled by an engine speed setting parameter, another orifice area 
downstream is scheduled by an engine pressure parameter and the third 
orifice area downstream is scheduled by an engine temperature parameter. 
The speed controlled orifice would be the effective orifice area for the 
regulator provided the speed schedule did not exceed either the engine 
pressure limit schedule or the engine temperature limit schedule. If 
subsequent engine power demands at the set speed condition causes either 
the pressure limit or temperature limit to be reached, the applicable 
parameter controlled orifice area becomes the effective area for the 
regulator and would control fuel flow to the engine as required to 
maintain operation at the scheduled limit condition for the particular 
parameter. 
SINGLE STAGE REGULATORS 
Many industrial process applications do not require the accuracy and wide 
range operating capability of the two stage servo operated regulators 
discussed above. The servo valve assembly configurations utilized for the 
aformentioned two stage regulators are in themselves single stage 
regulators which could be easily and economically adapted to satisfy such 
process applications. These single stage regulators provide a significant 
performance improvement over current available single stage regulators 
because of the use of the low force, low spring rate for pressure setting 
requirements. This is particularly true and evident for very high pressure 
applications in the area of 5000 psig or 10,000 psig range. 
BACK PRESSURE REGULATOR--SINGLE STAGE 
In FIG. 5 is described a single stage back pressure regulator adapted from 
the servo valve configuration described in FIG. 1. This regulator will 
control system pressure upstream of the regulator and is applicable for 
processes utilizing fluid flow at applicable "constant" pressure levels at 
flow rates less than the capacity of the positive displacement type pumps 
used in the system. This regulator would be ideally applicable for 
controlling the pressure between the pump and the nozzle of a high 
pressure water spray or drill. Water flow at 10,000 psig upstream of a 
nozzle will drill through rock and marble. At pressures between 3000 psig 
and 5000 psig, water spray will strip paint from metal sheeting or will 
remove a finite thickness of brick or granite facing as required to clean 
building exteriors. These processes do not require the performance 
capability of the two stage regulators previously discussed. Nozzles in 
these systems are usually incorporated at the end of a hand held flexible 
hose equipped with a hand operated valve to shut off fluid flow through 
the nozzle when necessary. 
Consider that the fluid pressure source between the pump and spray nozzle 
of such a water spray system is suitably connected to inlet passageway 412 
of the regulator shown in FIG. 5. The excess fluid flow not utilized by 
the nozzle spray is transported to contoured throttling port 417 at the 
end of passageway 412. Passageway 414 interconnects passageway 412 with 
pressure sensing chamber 456 formed between actuators 430 and 440 
respectively. Fluid flow passes through the throttling orifice formed 
between the opened edge of port 417 and metering edge 446(a) of piston 
land 446 on actuator 440 and flows through chamber 458 into exit 
passageway 416. A suitable piping connection would be used to transport 
the excess flow to the external system's vented supply tank. Passageways 
418 and 419 interconnect exit passageway 416 with chambers 452 and 454 
formed between respective piston lands 432 and 434 on actuator 430 and 
between piston lands 446 and 448 on actuator 440. These chambers, as 
previously noted for similar applications, intercept possible leakage from 
the high pressure sensing chamber 456 and from throttle port 417 and 
prevent the leakage from entering into the reference pressure sensing 
chambers 457 and 459 respectively. Chambers 457 and 459 are interconnected 
by passageways 421 and 423 and terminate at reference pressure port 425. A 
suitable piping connection can be used to connect an ambient pressure 
source to port 425. A connection in the ullage portion of the systems 
vented supply tank or even several inches below the liquid level in the 
tank could be used for this purpose. The force developed by compressing 
spring 402 provides the means for setting the regulators operating 
pressure level. 
As an example of the operating capabilities of a single stage back pressure 
regulator according to the above description for use in a 5000 psig water 
spray application, let the following be assumed: 
Dia. of actuator 440=D.sub.1 =0.6250 inches 
Bore dia. of actuator 440=D.sub.1B =0.6251 inches 
Dia. of actuator 430=D.sub.2 =0.6229 inches 
Bore dia. of actuator 430=D.sub.2B =0.6230 inches 
Dia. of rod 435=D.sub.3 =0.125 inches 
Spring rate of Spring 402=SR=10 lbs. per inch 
Water spray system pump capacity=6200 PPH (lbs. per hour) 
Regulator setting=5000 psig system pressure level when delivering flow 
through spray nozzle at 5200 PPH (approx. 10.4 gpm) 
To simplify the quantitative analysis, it will also be assumed: 
1. The rate of fluid flow through the regulator is in accordance with 
equation: 
EQU W.sub.f =12,000A.sub.o .sqroot..DELTA.P 
Where: 
W.sub.f =flow rate in PPH 
A.sub.o =Throttle port open area in square inches 
.DELTA.P=pressure loss across throttle port opening in lbs per square 
inches 
2. The fluid pressure loss through chamber 456, exit passageway 416 and 
connecting piping to the system's ambient vented supply tank is negligible 
for all flow conditions. 
3. Ambient atmospheric pressure from the vented supply tank is supplied to 
port 425 and to chambers 457 and 459. 
When 5200 PPH flow is expelled through the spray nozzle, the regulator must 
by-pass 1000 PPH flow (6200-5200 PPH) to the system tank. The initial 
opening area through throttle port 417 must be: 
##EQU2## 
Assume that the contoured throttle port 417 is comprised of four 
interconnected rectangular slots equally spaced about the circumference of 
the bore in chamber 458 and that the total equivalent width of the four 
slots is equal to one third of the circumference of the chamber bore and 
that the length of the slots are 0.437 inches long along the horizontal 
centerline of the assembly. The initial length of the port 417 slot 
openings when 1000 PPH is flowing is therefore: 
##EQU3## 
The initial spring force imposed by spring 402 to develop the 5000 psig 
pressure setting is calculated from an equation similar to that previously 
defined in the servo valve discussion section. 
EQU F.sub.S1 =P.sub.1 (A.sub.4 -A.sub.3)-Pr(A.sub.7 -A.sub.1) 
Considering: 
P.sub.1 =5000 psig 
Pr=0 psig 
A.sub.4 =(.pi./4)(D.sub.2.sup.2 -D.sub.3.sup.2) 
A.sub.3 =(.pi./4)(D.sub.2.sup.2 -D.sub.3.sup.2) 
Combining and rearranging: 
EQU F.sub.S1 =5000(.pi./4)(D.sub.1.sup.2 -D.sub.2.sup.2) 
Based on previously defined D.sub.1 and D.sub.2 values 
##EQU4## 
The initial spring compression necessary to develop this force is: 
##EQU5## 
The regulator must accomodate and by-pass the full system pump capacity of 
6200 PPH flow to the tank when flow through the nozzle is stopped. The 
operating conditions of the regulator can be established for this new flow 
condition by simultaneously solving the following equations: 
(Note, subscript 2 defines the applicable parameters at the new flow 
conditions.) 
##EQU6## 
Simultaneously solving the above equations results in the following: 
.DELTA.P.sub.2 =5045.2 psig 
A.sub.02 =0.007274 in.sup.2 
L.sub.2 =0.01111 inches 
F.sub.S2 =10.3841 lbs. 
The regulator would only develop an increase of less than 1% in pressure 
setting (from 5000 psig to 5045.2 psig) when flow through increases from 
1000 PPH to 6200 PPH. This is a significant improvement over present 
available single stage regulators for the noted pressure setting and flow 
range operating conditions. In fact, an expensive two stage regulator of 
conventional design would be required to approach this "droop" 
characteristic of less than 1% for the same conditions. 
It should also be noted that when the throttle port 417 is opened for the 
full 0.437 inch length, the pressure loss through the port at 6200 PPH 
flow would be 
##EQU7## 
The regulator would then have a pressure setting adjustment range between 
4 psig to over 5045 psig provided the spring adjustment means (which 
compresses the spring) has a working range in excess of: 
##EQU8## 
It should be noted that the servo valve assembly shown in FIG. 1 would be 
identical in configuration with the single stage back pressure regulator 
if the following changes were made: (Ref. FIG. 1.) 
1. Relief port 16 and passageway 51 are eliminated. 
2. Passageway 88 relocated to interconnect passageway 89 with passageway 53 
and is designated inlet passageway. 
3. Servo port 14 enlarged and relocated to be throttled by piston metering 
edge 44b. 
Physically the regulator in FIG. 5 would be longer and larger to accomodate 
the much larger flowrates through the passageways and the much larger 
throttle port. 
SINGLE STAGE PRESSURE AND FLOW REGULATORS 
It should be evident that, in like manner, single stage pressure and flow 
regulators could be created by adapting any servo valve configuration 
described herein for the two stage regulators. In FIG. 6 is described a 
single stage flow regulator based on an adaptation of the servo valve 
assembly in FIG. 3. Note that fluid pressure in passageway 816 upstream of 
orifice 860 is also applied in chamber 856 formed between actuators 830 
and 840. The pressure in outlet passageway 818 downstream of orifice 860 
is interconnected with chambers 857 and 859 by passageways 821 and 823 
respectively. Piston land 830 will vary the opening of port 817 located at 
the end of inlet passageway 812 to maintain a set pressure loss across 
orifice 860 as established by the compression setting of spring 802. 
In FIG. 7 is described a single stage pressure regulator adapted from the 
servo valve assembly in FIG. 2. Chambers 957 and 959 are connected to 
regulated pressure outlet 919 by passageways 921 and 923 respectively. A 
desired reference pressure source is suitably connected to passageway 925 
and is applied into chamber 956. This regulator will position piston land 
932 to throttle fluid flow passing through port 917 at the end of inlet 
passageway 912 to maintain a fluid pressure level in outlet 919 as set by 
the compression of spring 902. 
These single stage regulators will have a "droop" characteristic and the 
set flow or pressure level will not be as accurate as that attainable with 
the two stage regulators previously described. They will, however, be more 
economical for applications not requiring such accuracy or range 
capability. 
SUMMARY OF ADVANTAGES OF THE INVENTION 
1. Single unitized servo valve assembly. 
2. Highly Accurate--low force, low rate springs used for pressure sensing 
applications for any pressure level, no seals used on servo or throttle 
valves. 
3. Unlimited Pressure Level Application--no diaphragms, bellows or thin 
sectioned membranes used. 
4. Very Responsive--regulation very responsive due to low spring rate 
(force/deflection) of spring used in servo valve assembly. 
5. Highly Reliable--Design uses no thin section membranes, bellows, etc., 
that are subject to high fatigue type stresses and which are usually the 
cause of most regulator failures.