Patent Publication Number: US-7582009-B1

Title: Adjustable air volume regulator for heating, ventilating and air conditioning systems

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
   This application claims the benefit of PPA Ser. No. 60/651,361 filed Feb. 10, 2005 

   FEDERALLY SPONSORED RESEARCH 
   Not applicable 
   SEQUENCE LISTING OF PROGRAM 
   Not applicable 
   BACKGROUND OF THE INVENTION  
   1. Field of Invention 
   This invention relates to an air volume regulator, more specifically to an improved “airflow powered” air volume regulator as used to control the flow of conditioned air in the ductwork of heating, ventilating and air conditioning (HVAC) systems, clean rooms or fume hood systems. 
   2. Background of the Invention 
   In HVAC systems, air is supplied from a central air conditioning system to several outlet devices such as grilles and diffusers in the rooms or spaces being conditioned. Once a HVAC system is installed, the airflow through the ductwork system must be adjusted or balanced. This insures that each room or space obtains the specified volumes of conditioned air from the central system. In its simplest form, this can be done by using manually adjustable dampers. They are placed within the supply air and return air ductwork to reduce the airflow in areas where it exceeds the specified amount. There is an inherent problem with this method. When one damper is adjusted, the pressure level throughout the ductwork system will change. Any change in the ductwork system pressure will affect the flow or air past every other damper including the previously adjusted dampers. On large systems, it quickly becomes impractical to attempt to balance the ductwork system using dampers alone. To solve this problem, air volume regulators are added. They are designed to limit the supply of conditioned air to the desired amount and this, irregardless of the pressure at their inlet. Also, once calibrated, the airflow is not affected by subsequent variations in system pressure. The accepted industry standard airflow variation is +/−5% of the specified airflow volume (4.7 L/s below 94 L/s or 10 cfm below 200 cfm) over the airflow regulator&#39;s pressure range. 
   Furthermore, heating, cooling and ventilating loads in a room or space vary in time. It has become common practice to stabilize the temperature of the rooms or spaces by:
         (a) varying the volume of conditioned air supplied to each room or space or   (b) using heating coils downstream from the air volume regulator to heat the volume of cool conditioned air being supplied to the room or space or   (c) using a dual ductwork layout for the HVAC system known as dual-duct system: one supplying hot air, the second supplying cool air and a mixing valve upstream of the air volume regulator or   (d) using combinations of the above.       

   Air volume regulators fall into one of two general groups based on the source of energy that is used to drive them: “airflow powered” and “externally powered”. 
   “Airflow powered” air volume regulators function using the energy of the air flowing in the ductwork system. This source of energy is in the form of air static pressure and air velocity pressure (called dynamic pressure). The scope of this invention is limited to improved “airflow powered” air volume regulators. 
   Briefly, “externally powered” air volume regulators operate using an external energy source such as pneumatic pressure or electricity. They require an airflow sensor, a signal amplifier, an actuator and an adjustable airflow restricting device or damper to regulate the flow of air. 
   As mentioned above, the energy source used to drive the “airflow powered” air volume regulator comes from two types of pressure present in HVAC systems: static pressure and dynamic pressure. The static pressure induces the movement of the air through the ductwork towards the outlets while the dynamic or velocity pressure is generated by the movement of the air at a given point within the ductwork. The higher the air velocity the greater the dynamic pressure. 
   The flow of air through the ductwork of an HVAC system is governed by the following basic formulas: 
   No. 1: The relationship between air velocity and dynamic pressure is given by the following
 
 P   d =Dynamic Pressure=Constant×( V   a ) 2  
 
   No. 2: The sum of the static pressure and dynamic pressure is called the total pressure:
 
 P   t   =P   s   +P   d  
 
   No. 3: The airflow rate is equal to the air velocity times the area of the duct cross section.
 
 Q   a   =V   a   ×A   duct  
         where Q a =Air volume rate   V a =air velocity   A duct =area of the duct cross section       

   A conclusion of formula no. 2 is that, under the idealized conditions of constant total pressure (i.e. no losses due to friction and turbulence), the static pressure and dynamic pressure can be converted from one to the other. A decrease of one entails an equal increase of the other. 
   A conclusion of formula no. 3 is that, for a constant volume flow, an increase in the duct cross section entails a proportional decrease in the air velocity. Conversely, a decrease in the duct cross-section entails a proportional increase in the air velocity. 
   Combining the 3 formulas, we can conclude that, for a constant total pressure (i.e. no losses due to friction and turbulence):
         an increase in the duct cross section means
           a decrease in dynamic pressure (associated with a decrease in air velocity)   an equal increase in static pressure. This is known as “static regain”.   
           a decrease in the duct cross section means
           an increase in dynamic pressure (again associated to the air velocity) and   an equal decrease in static pressure.   
               

   3. Background of the Invention—Discussion of Prior Art 
   A—INTERACTION WITH THE HVAC SYSTEM 
   As stated above, an air volume regulator is required when the airflow rate in the ductwork system exceed the desired amount. This happens when more static pressure is present in the air duct than is required to move the air to the outlet devices. Although all air volume regulators inherently create a pressure loss due to air friction and air turbulence as the air flows through them, prior art “airflow powered” air volume regulator also requires some amount of pressure to drive the flow control means. The sum of these pressure losses is called the regulator minimum static pressure. Under most conditions, this extra control pressure required to drive the flow control means is of little consequence. Excess pressure is usually present at the inlet of the air volume regulator. In the cases where the air volume regulator is installed in a area with little or no excess static pressure (i.e. at the far limits of the central air distribution system), the control pressure requirement may be greater than the available excess pressure. Consequently, the air volume regulator will be incapable of controlling the airflow and the desired airflow rate will not be attained. Thus a desirable characteristics of an air volume regulator is that it have a negligible pressure loss due to air friction and turbulence and more important still, require virtually no static pressure to bring the flow of air under control. Although many forms of prior art have been proposed, none have met this challenge. 
   B—OPERATION 
   An air volume regulator functions by varying its internal airflow passageway(s) so as to maintain a substantially constant airflow rate. In “airflow powered” air volume regulators, the air pressure and velocity drive some form of airflow restricting means. A spring is then used to counterbalance the forces acting on the restricting means (called the counterbalance spring). 
   The graph in  FIG. 1  shows, for a constant flow of air, the variation in the cross sectional area of the narrowest portion or throat of the airflow passageway as the static pressure differential across the air volume regulator increases from 25 pascals to 1000 pascals (0.1″ to 4″ w.g). 
   The graph in  FIG. 1  is derived from the following:
         (a) with the throat open to its maximum, the static pressure drop is 25 pascals (0.1″ w.g.); this pressure is mainly to drive the flow control means and frictional loses due to turbulence,   (b) the variation in the area of the throat is inversely proportional to the air velocity at the throat, i.e. if the velocity goes up, the area must go down proportionally, a corollary of formula 3 above,   (c) the static pressure differential or static pressure drop between the inlet and outlet of the air passageway is substantially converted to dynamic pressure in the throat of the passageway. This assumes that the total pressure remains substantially constant (negligible loses due to friction or turbulence)   (d) in the throat, the static pressure drop is proportional to the square of velocity at that point (the drop in static pressure is totally converted to dynamic pressure) i.e. if the static pressure drop doubles, the velocity in the throat will quadruple.       

   Combining the three previous statements, an exponential equation is obtained: 
   
     
       
         
           
             Area 
           
           = 
           
             
               constant 
             
             
               
                 static  pressure  drop 
               
             
           
         
       
     
       
       
         
           where the constant depends on the air volume regulator design. 
         
       
     
  
     FIG. 1  is a graph of this exponential equation. The cross-sectional area of the throat must drop quickly as the pressure differential rises. It then starts to level off to a point were very little reduction of the throat is required to control the airflow (approximately 500 pascals or 2″ w.g.). Also, as the pressure differential rises from 25 pascals to 125 pascals (0.1″ to 0.5″ w.g.), the area of the throat must be reduced by over 50%. Furthermore, since the air volume regulator contains moving parts, mechanical friction is present and this inhibits the reduction of the throat. Thus to operate reliably, the airflow regulator must be capable of initiating and maintaining control of the airflow using and/or amplifying the very low forces generated by these low pressures. I have found no prior art that proposes an “airflow powered” air volume regulator with this latter capability. 
   The sum of all mechanical friction in an airflow regulator generates an adverse effect on its operation. This can be visualized by drawing an hysteresis graph: a graph is plotted of the airflow rate versus inlet pressure as the pressure is slowly increased up to the upper limit of the airflow regulator&#39;s pressure range then, on the same graph, is plotted the airflow rate versus inlet pressure as the pressure is slowly decreased down to the lower limit of the pressure range. The two curves do not coincide with one another. The reason for this is that as the pressure increases, the air velocity increases and the airflow regulator tends to reduce the airflow passageways so as to maintain the specified airflow rate. But mechanical friction within it tends to resist this reduction and the correct passageway size is not attained. The airflow passageways are consequently a little to wide and thus the airflow rate will be slightly above the specified airflow rate. Conversely, as the pressure decreases from the upper limit of the pressure range, the air velocity decreases and the airflow regulator tends to increase the airflow passageways so as to maintain the desired airflow rate. Again mechanical friction within it tends to resist this increase. The airflow passageways are consequently a little to narrow and thus the airflow rate will be slightly below the specified airflow rate. The more mechanical friction is found in an airflow regulator, the greater the difference between the specified airflow rate and the actual airflow rate. As outlined above, this difference must not exceed 5% of the specified airflow rate. 
   In practice, attempts made in commercially available prior art to reduce the regulator minimum static pressure below 100 pascals (0.3″ w.g.) for “airflow powered” units have been unsuccessful. In referring to U.S. Patents such as 3,049,146 to Hayes (1962), 2,890,716 to Werder (1959), 3,338,265 to Kennedy (1967) and 4,009,826 to Walker (1977), forcing the airflow to pass through perforated screens has a particularly adverse effect on the regulator minimum static pressure (this generates high airflow friction and turbulence losses). As mentioned above, particular attention must be paid to mechanical friction in the moving parts of the air volume regulator. All prior art embodiments using components sliding on shafts encounter, over time, binding of some kind when dirt particles get lodged in the sliding bearing. This is the case with U.S. Pat. Nos. 3,204,664 to Gorchev et al. (1965), 3,763,884 to Grassi et al. (1973), 3,958,605 to Nishizu et al. (1976) and 4,009,826 to Walker (1977). Adding lubricant does not improve this inconvenience since again, over time, the lubricant attracts dirt particles to form a abrasive paste that resists movement. 
   C—COUNTERBALANCE SPRING 
   The air volume regulator should include a means to easily vary the desired airflow rate&#39;s setpoint at will once installed in the ductwork system. This is made necessary by changing conditions such as occupancy of the areas, insolation and outside temperature. It is common practice to add an optional actuator to adjust the setpoint mechanism and control it remotely. As noted above, a spring is used to counterbalance the forces acting on the flow restricting means. As stated in U.S. Pat. Nos. 4,306,585 to Manos (1981), 4,009,826 to Walker (1977), 3,958,605 to Nishizu (1976), 3,942,552 and 3,939,868 to Logsdon (1976), 3,763,884 to Grassi et al. (1973), 3,565,105 to Murakami (1971) and 3,037,528 to Baars et al. (1962), varying the initial load of the counterbalance spring by adjusting the initial spring deflection is only effective for small variations in the airflow setpoint. As stated in these patents, the spring quickly become either to stiff or to soft and the air volume regulator ceases to adequately control the airflow. Although not stated in the following patents, I also believe this to be true for U.S. Pat. Nos. 4,633,900 to Suzuki (1987), 3,967,642 to Logsdon (1976), 3,433,410 to Warren (1969), the embodiment in  FIG. 7  of 3,276,480 to Kennedy (1966), 3,204,664 to Gorchev et al. (1965), 3,049,146 to Hayes (1962) and 2,890,716 to Werder (1959). Thus for a particular air volume regulator at a given flow rate corresponds a spring stiffness known as its spring rate or spring constant and is defined as the force generated divided by the spring deflection. In general terms, the spring rate is the “force-displacement” characteristic of the spring. To vary the airflow setpoint requires that the spring stiffness be varied or several springs be used over the operating range of the regulator. As shown in U.S. Pat. Nos. 4,009,826 to Walker (1977), 3,939,868 (1976) and 3,942,552 (1976) to Logsdon and my own U.S. Pat. No. 4,130,132 (1978), relatively complex mechanisms are proposed to vary the spring stiffness. 
   D—ACTUATORS 
   As mentioned above, an actuator may be used to action the flow rate setpoint mechanism. The control signal to the actuator is generally supplied by a thermostat. Actuators may be either pneumatic or electric driven. But pneumatic actuators can be a problem when the flow restricting forces applied to the counterbalance spring are also carried by the actuator. Pneumatic actuators have an inherent load dependant stroke or travel due to the compressibility of the air pushing the actuator&#39;s piston. Since the flow restricting forces vary in time due to changes in the static pressure upstream from the air volume regulator, so does the load on the counterbalance spring and thus the actuator. The pneumatic actuator&#39;s piston will move or slip under the varying load with the ensuing unjustified change in the flow rate setpoint. This phenomena is clearly outlined in the report “Factors that work to defeat the application of the “spring and cone” type valves in laboratory and other precision airflow systems” by Swiki A. Anderson, Ph.D., P. E., (Swiki Anderson &amp; Associates, Inc. 1516 Shiloh Avenue, Bryan, Tex. 77803). The thermostat will sense a variation in the temperature of the room caused by the change in the flow rate and adjust the pressure to the pneumatic actuator to rectify the unjustified change and its ensuing discomfort to the occupants. This is the case in U.S. Pat. Nos. 4,633,900 to Suzuki (1987), 4,175,583 to Finkelstein et al. (1979), 3,958,605 to Nishizu et al. (1976), 3,942,552 to Logsdon (1976), 3,204,664 to Gorchev (1965) and my own U.S. Pat. No. 4,130,132 (1978). 
   Further concerns involving the actuator are:
         To facilitate field servicing and repairs, the actuator should not be situated inside the air volume regulator or its housing such as U.S. Pat. Nos. 3,976,244, 3,942,552 and 3,939,868 to Logsdon (1976) and my own U.S. Pat. No. 4,130,132 (1978)   Its replacement should not affect the calibration of the air volume regulator such as my own U.S. Pat. No. 4,130,132 (1978).
 
E—ZERO FLOW
       

   When an actuator is used to vary the flow rate set point and under certain conditions, it is common practice in HVAC systems to restrict the flow completely (substantially zero flow is the accepted industry standard leakage of 2% of the maximum airflow capacity at the maximum operating pressure of the regulator). The flow restricting means must then be able to block the flow of air through the air volume regulator. In prior art, this is not possible with U.S. Pat. No. 3,958,605 to Nishizu et al. (1976), U.S. Pat. No. 4,009,826 to Walker (1977) or U.S. Pat. No. 4,633,900 to Suzuki (1987) because of leakage at edges of the flow restricting plates. Furthermore, this is not possible with U.S. Pat. Nos. 3,942,552 or 3,939,868 to Logsdon (1976) because the counterbalance spring never totally releases the flow restricting means or with U.S. Pat. No. 4,009,826 to Walker (1977) and U.S. Pat. No. 3,204,664 to Gorchev (1965) because of leakage at the edges of their sliding flow restrictors. 
   F—PULSATION 
   A phenomena that is well known in prior art and unique to “airflow powered” air volume regulators is their propensity to flutter, oscillate or pulsate when the air stream at their inlet is unstable. This inherent characteristic is due to the use of a spring to counterbalance the airflow restricting forces within the air volume regulator. Variations in the pressure upstream from the air volume regulator caused by turbulence or other instabilities can induce pressure pulses that travel down the ductwork to the air volume regulator. These fluctuations induce a rapid rise and fall in pressure usually lasting less than a second. If the amplitude of the pressure pulse is significant, the air volume regulator will react rapidly to constrict the airflow passage on sensing the rise in pressure then open the airflow passage on sensing the drop in pressure. But the inertia of the apparatus is such that the air volume regulator will tend to be out of phase with the quick change in pressure: over-constricting the airflow passage as the pressure starts to return to normal or under-constricting the airflow passage once the pressure has returned to its initial level. This out of phased reaction sets in motion the pulsation, as the spring-inertia combination oscillates between extremes driven by the energy of the air upstream of the apparatus as a car with defective shock absorbers when it hits a bump in the road. Dampening means must then be included to brake the cycle. 
   In prior art, U.S. Pat. No. 3,276,480 Kennedy (1966) and U.S. Pat. No. 3,763,884 to Grassi et al. (1973) employ dashpots and U.S. Pat. No. 3,049,146 to Hayes proposes a wear plate but their inherent friction hinders the airflow tracking of the air volume regulator and creates an unacceptably large hysteresis in its control. It is to be noted again that all mechanical friction within the apparatus prevents it from operating at low pressures. U.S. Pat. No. 3,204,664 to Gorchev (1965) teaches an air bellows with a flow orifice but the entrapped air is compressible and acts like a spring (air spring). The addition of mass to create inertia such as flywheels is shown in U.S. Pat. No. 3,060,960 to Waterfill (1962), but this method only lowers the natural frequency of the spring-mass combination: pulsation can still occur but at a lower frequency. The only true dampening that will dissipate the energy is due to the mechanical friction of this device. Under certain condition, I have found that the addition of inertia alone is ineffective. 
   In summary, the major drawbacks in prior art “airflow powered” air volume regulators are the following:
         The minimum static pressure required by the air volume regulator to start controlling the airflow rate is relatively high. The long-felt need for an air volume regulator functioning reliably at pressures at or below 25 pascals (0.1″ w.g.) is unsolved.   The means for varying the airflow rate setpoint remain relatively complex, ineffective or in some cases, none existent. In some prior art embodiments, the flow rate set point may “slide” when a pneumatic actuator is employed.   Most prior art embodiments cannot attain “zero flow” when an actuator is proposed to vary their flow rate setpoint.   They have a propensity to flutter, oscillate or pulsate when the airstream at their inlet is unstable. Dampening means are proposed but either generate excessive mechanical friction or, under certain conditions, are ineffective.       

   BACKGROUND OF THE INVENTION—OBJECTS AND ADVANTAGES 
   Accordingly, several objects and advantages of my invention are:
         (a) To provide an “airflow powered” air volume regulator that reliably controls the airflow to within the “HVAC Industry Standard Variation” of +/−5% of specified airflow rate (or 4.7 L/s below 94 L/s) (or 10 cfm below 200 cfm) over its full airflow range.   (b) To provide an “airflow powered” air volume regulator that solves the long-felt need to initiate the control of the airflow at a pressure of 25 pascals (0.1″ w.g.) or less over its full airflow range.   (c) To provide an air volume regulator requiring only one counterbalance spring to cover its full operating range.   (d) To provide an air volume regulator whose setpoint will not “slide” when a pneumatic actuator is included to vary its airflow setpoint.   (e) To provide an air volume regulator that can substantially shut-off the airflow when an actuator is installed to vary its airflow setpoint.   (f) To provide an air volume regulator that will substantially control its propensity to flutter, oscillate or pulsate when the air stream at its inlet is unstable.       

   Other objects and advantages are:
         (g) To provide an air volume regulator which, when combine with an optional actuator means to vary its airflow setpoint, requires at most 4 Nm (35 lbs-in) of torque with airflow rates as high as 944 L/s (2000 cfm).   (h) To provide an air volume regulator with few moving parts and substantially no mechanical friction between them.   (i) To provide an air volume regulator with its counterbalance spring and associated linkage removed from the airstream to eliminate the possibility of dirt particles within the airstream lodging in the pivot points of the moving parts and creating undesirable friction. This also permits servicing of the unit without shutting down the supply fan.   (j) To provide an air volume regulator which permits field adjustment from the exterior of the unit. When an actuator is required to vary its airflow setpoint, it is situated on the exterior of the air volume regulator and can be easily added or replaced in the field without requiring modifications or without affecting the air volume regulator&#39;s calibration.       

   Still further objects and advantages of my invention will become apparent from a consideration of the ensuing description and drawings. 
   SUMMARY 
   In accordance with the present invention, an air volume regulator that will maintain the flow of air moving through it substantially constant at inlet static pressures of 25 pascals (0.1″ w.g.) or less. It comprises a pair of gates that swing into the airflow, a counterbalance spring assembly with an adjustable spring rate and a cable driven “limited torque” flywheel. 

   
     DRAWINGS—FIGURES 
       FIG. 1  shows a graph of the variation of the cross-sectional area of the throat of the airflow passageway versus the static pressure differential between the inlet and throat for a constant flow of air. 
       FIG. 2  shows a graph of the variation of the force-displacement characteristic versus the angle of incidence of the counterbalance spring as illustrated in  FIG. 3   a  through  3   c.    
       FIG. 3   a  through  3   c  show the relationships between the input and output variables as the counterbalance spring is rotated between 0 and 90 degrees. 
       FIG. 3   d  through  3   g  show how the spring initial tension can be varied as the adjustable spring arm rotates. 
       FIG. 4   a  is an isometric view of the preferred embodiment showing the airflow section of the air volume regulator with part of the exterior shell broken away and the casing of spring counterbalance system outlined in the foreground. 
       FIG. 4   b  is an isometric view of the preferred embodiment showing the spring counterbalance system with its protective shroud removed and the airflow section beyond. 
       FIG. 4   c  is a partial exploded isometric view of the preferred embodiment showing the proposed pivot hinge. 
       FIG. 4   d  is a partial isometric view of the preferred embodiment showing the proposed pivot hinge assembly. 
       FIG. 4   e  is an isometric view of the airflow setting quadrant of the preferred embodiment. 
       FIG. 4   f  shows an exploded view of the dampener flywheel assembly of the preferred embodiment. 
       FIG. 4   g  shows a typical cross-section through the dampener flywheel and the pivot bearings. 
       FIGS. 4   h  and  4   i  show the air volume regulator viewed from below with and without an option actuator. 
       FIGS. 5   a,    5   b,    5   c ,  5   d  and  5   e  are sectional views of the preferred embodiment showing the positions assumed by the components under various conditions of upstream pressure and airflow settings where:
           FIG. 5   a  shows a cut-away plan view of the air volume regulator at maximum flow rate and at minimum pressure.     FIG. 5   b  shows a cut-away plan view of the air volume regulator at maximum flow rate and at maximum pressure.     FIG. 5   c  shows a cut-away plan view of the air volume regulator at a reduced flow rate and at minimum pressure.     FIG. 5   d  shows a cut-away plan view of the air volume regulator at a reduced flow rate and at maximum pressure.     FIG. 5   e  shows a cut-away plan view of the air volume regulator at zero flow rate.       
       FIG. 6   a  is an isometric view of an alternative embodiment of the airflow section of the air volume regulator with part of the exterior shell broken away and the protective shroud of the spring counterbalance system outlined in the foreground. 
       FIG. 6   b  shows a cut-away plan view of the alternative embodiment of  FIG. 6   a  at maximum flow rate and at minimum pressure. 
       FIGS. 7   a  and  7   b  show cut-away plan views of alternatives to the coupling cable and cable cams with the air volume regulator at maximum flow rate and no air entering the unit. 
       FIG. 8   a  shows the upgrading of an existing air volume regulator design with an adjustable counterbalance spring assembly and a cable driven flywheel. 
       FIGS. 8   b  and  8   c  show the details of the configuration of the drive cables of  FIG. 8   a.    
       FIG. 9  show a cut-away plan view of an alternative to the eccentric cam and guide angle. 
   

   DRAWINGS—LIST OF REFERENCE NUMERALS 
   In the drawings, related elements of a given part have the same number but different alphabetic suffixes. 
   
     
       
         
             
             
             
             
           
             
                 
             
           
          
             
               10 
               inlet section 
               11 
               optional transition section 
             
             
               11a 
               included angle of transition section 
               12 
               flow constricting section 
             
             
               12a 
               sidewall adjacent to idler gate 
               12b 
               sidewall adjacent to drive gate 
             
             
               12c 
               plenum adjacent to idler gate 
               12d 
               plenum adjacent to drive gate 
             
             
               13 
               outlet section 
               13a 
               sidewall of outlet section 
             
             
               13b 
               sidewall of outlet section 
               13c 
               opening in sidewall of outlet section 
             
             
               14 
               thermal &amp; acoustic material 
               15 
               idler gate 
             
             
               15a 
               idler gate upstream end 
               15b 
               idler gate pivot tab 
             
             
               15c 
               idler gate locking tab (not shown) 
               15d 
               idler gate bracket 
             
             
               16 
               drive gate 
               16a 
               drive gate upstream end 
             
             
               16b 
               drive gate pivot tab 
               16c 
               drive gate locking tab 
             
             
               16d 
               drive gate bracket 
               17 
               entering air stream 
             
             
               17a 
               air stream passing idler gate 
               17b 
               air stream passing drive gate 
             
             
               17c 
               air stream at outlet 
               17d 
               air stream between the gates 
             
             
               18 
               gate lever 
               18a 
               gate lever link A 
             
             
               18b 
               gate lever link B 
               18c 
               gate lever link C 
             
             
               18d 
               gate lever link D 
               19 
               “V” baffle 
             
             
               19a 
               “V” baffle openings 
               19b 
               “V” baffle curved air guide 
             
             
               19c 
               “V” baffle pivot tab slot 
               19d 
               “V” baffle arm 
             
             
               19e 
               “V” baffle arm 
               20 
               coupling pin 
             
             
               21 
               coupling pin slot 
               22 
               gate lever fasteners 
             
             
               23 
               follower bearing 
               24 
               track 
             
             
               25 
               chassis 
               25a 
               chassis guide plate 
             
             
               25b 
               chassis opening 
               25c 
               chassis access opening 
             
             
               25d 
               chassis end panel 
               25e 
               shuttle pivot slots 
             
             
               26 
               removable shroud 
               27 
               shroud fasteners 
             
             
               28 
               shuttle 
               28a 
               shuttle pivot tabs 
             
             
               28b 
               counterbalance spring pivot hole 
               29 
               cam 
             
             
               30 
               counterbalance spring 
               31 
               spring arm 
             
             
               32 
               actuator shaft 
               33 
               low friction sleeve bearing 
             
             
               34 
               threaded pivot bolt 
               35 
               threaded pivot bolt locking nut 
             
             
               36 
               extension nut 
               37 
               threaded eye bolt 
             
             
               38 
               adjusting nut 
               39 
               guide angle 
             
             
               39a 
               upstream guide angle arm 
               39b 
               downstream guide angle arm 
             
             
               40 
               eccentric circular cam 
               41 
               connecting link 
             
             
               42 
               retaining angle 
               43 
               flywheel drive bow 
             
             
               43a 
               flywheel drive bow cable hook 
               44 
               drive cable tensioning spring 
             
             
               45 
               flywheel drive cable 
               46 
               flywheel drive bushing 
             
             
               47 
               dampener flywheel assembly 
               48 
               conical cup bearing 
             
             
               48a 
               conical cup bearing recess 
               49 
               conical cup bearing nut 
             
             
               50 
               internal tooth retaining ring 
               51 
               spring alignment shoulder washer 
             
             
               52 
               compression spring 
               53 
               flywheel disk 
             
             
               54 
               flywheel pivot pins 
               54a 
               flywheel pivot pins pointed end 
             
             
               54b 
               flywheel pivot pins blunt end 
               55 
               flywheel shaft 
             
             
               55a 
               guide portion of flywheel shaft 
               55b 
               flywheel support flange 
             
             
               56 
               cam fastener 
               57 
               flywheel bow fastener 
             
             
               58 
               guide angle fastener 
               59 
               chassis sliding seal 
             
             
               60 
               airflow setting quadrant assembly 
               61 
               flow indicator arm 
             
             
               61a 
               pointed end of flow indicator arm 
               62 
               “U” shaped adjustable limit stops 
             
             
               62a 
               limit stops leg 
               62b 
               limit stops leg cusp 
             
             
               63 
               slotted quadrant 
               64 
               carriage bolt 
             
             
               65 
               wing nut 
               66 
               quadrant fastener 
             
             
               67 
               air volume scaled decal 
               68 
               pivot screw 
             
             
               69 
               shroud seal 
               70 
               optional actuator 
             
             
               71 
               coupling link 
               72 
               (not used) 
             
             
               73 
               coupling cable 
               73a 
               cable cam 
             
             
               73b 
               cable cam 
             
             
               74 to 109 
               (not used) 
             
             
               110 
               retaining arm 
               110a 
               retaining arm pivot pin 
             
             
               110b 
               spring pivot hole 
               111 
               tension link 
             
             
               111a 
               tension link pivot hole 
               111b 
               tension link hook 
             
             
               112 
               airframe crown 
               113 
               curtain frame pivot pin 
             
             
               114 
               curtain frame 
               114a 
               pivot point of drive bow 119 
             
             
               114b 
               pivot point of drive bow 120 
               114c 
               anchoring point of equalizer cable 
             
             
               114d 
               anchoring edge of flexible curtain 
               114e 
               curtain frame pivot arm 
             
             
               115 
               impervious flexible curtain 
               116 
               pervious pitched sidewall 
             
             
               117 
               impervious sidewall 
               118 
               equalizer cable 
             
             
               119 
               drive bow 
               119a 
               anchorage point of drive cable 
             
             
               119b 
               anchorage point of drive cable 
               120 
               drive bow with tightener 
             
             
               120a 
               tensioning link pivot point 
               120b 
               anchorage point of drive cable 
             
             
               120c 
               tensioning spring hook 
               121 
               tensioning link 
             
             
               121a 
               anchorage of drive cable 
               121b 
               anchorage of tensioning spring 
             
             
               122 
               impervious end wall 
               123 
               mounting plate 
             
             
               124a 
               flywheel drive cable 
               124b 
               flywheel drive cable 
             
             
               125 
               tensioning spring 
               126 
               inlet opening 
             
             
                 
             
          
         
       
     
   
   DETAILLED DESCRIPTION—PREFERRED EMBODIMENT—FIGS.  4   a  to  4   i    
   The preferred embodiment of the air volume regulator of the present invention is illustrated in  FIG. 4   a —Isometric view of the flow control module and  FIG. 4   b —Isometric view of the spring counterbalance system. Referring to  FIG. 4   a,  the air volume regulator is comprised of a housing of generally rectangular section having:
         (a) an upstream end or inlet  10 , generally cylindrical in shape, through which conditioned air under pressure is supplied to the air volume regulator,   (b) a flow constriction section  12  with sides walls  12   a  and  12   b,      (c) an optional expansion or transition section  11  required if the area of inlet  10  is smaller than that of flow constriction section  12 . The length of transition section  11  is sufficient to permit the efficient conversion of at least 75% of the reduction in dynamic pressure between inlet  10  and transition section  11  to static pressure as per the known practice of “static regain”.   (d) an outlet or downstream section  13  through which the conditioned air is conveyed to the room or space to be conditioned.       

   As is conventional, the inner surfaces of the walls of transition section  11 , flow constriction section  12 , and outlet section  13  may be covered with a suitable thermal acoustic insulating material  14 . Flow constriction section  12  includes two opposing pivoted sidewalls or gates  15  and  16  and a “V” shaped baffle  19 . Gate  15  is defined as an idler gate and gate  16  as a drive gate. Baffle  19  is positioned centrally in section  12  with its apex pointing upstream to divide entering airstream  17  into two airstreams  17   a  and  17   b.  Upstream ends  15   a  and  16   a  of gates  15  and  16  respectively are curved away from baffle  19  so as to direct airstreams  17   a  and  17   b  against baffle arms  19   d  and  19   e.  Idler gate  15  is attached to baffle arm  19   d  by two pivot tabs  15   b  at its downstream end such that it can freely swing between baffle arm  19   d  and sidewall  12   a.  Similarly, drive gate  16  is attached to baffle arm  19   e  by two pivot tabs  16   b  at its downstream end such that it can freely swing between baffle arm  19   e  and sidewall  12   b.    FIG. 4   c  shows a partial exploded isometric view of baffle arm  19   e  at its downstream end and drive gate  16  not installed;  FIG. 4   d  shows a partial isometric view of baffle arm  19   e  at its downstream end and viewed from the airstream side with drive gate  16  installed. Pivot tabs  16   b  are identical to pivot tabs  15   b  on gate  15 . Pivot tabs  16   b  are inserted in two pivot tab slots  19   c  in baffle arm  19   e  and two locking tabs  16   c  are compressed as they pass through pivot tab slots  19   c.  When pivot tabs  16   b  are fully inserted, locking tabs  16   c  snap back to their original shape and lock gate  16  in slots  19   c.  Pivot tabs  16   b  now act as a hinge. Pivot tab slots  19   c  are sufficiently wide and high so that gate  16  swing freely on pivot tabs  16   b  without being overly loose. 
   Now returning back to  FIG. 4   a,  the height of gates  15  and  16  is substantially the same as inlet  10  and their total length is approximately 3 times their height. Baffle  19  expands downstream from its apex to sides  12   a  and  12   b  respectively of flow constriction section  12 . It is fixedly sealed to flow constriction section  12  on all its outer edges. Two openings  19   a  are cut in baffle arms  19   d  and  19   e  to permit the free passage of the air towards outlet section  13 . The height of openings  19   a  is smaller than the gate height such that gates  15  and  16  can completely cover them. The length of openings  19   a  is at approximately 2/3 the length of baffle arms  19   d  and  19   e.  A curved baffle air guide  19   b  extends from the upstream edges of openings  19   a,  downstream between baffle arms  19   d  and  19   e.  Optionally, baffle air guide  19   b  may be ogival, elliptical or round extending downstream. 
   An idler gate bracket  15   d  and a drive gate bracket  16   d  are fixedly attached to gates  15  and  16  respectively extending into the airstream and through openings  19   a.  Gate brackets  15   d  and  16   d  movably link gates  15  and  16  together through a coupling pin  20  sliding in an alignment slot  21 . Coupling pin  20  forces gates  15  and  16  to operate in unison but in opposite directions with substantially equal angular rotations. A gate lever  18  is fixedly attached to the downstream edge of drive gate bracket  16   d  with two fasteners  22  and extends from it around the downstream end of baffle  19  and through an opening  13   c  in sidewall  13   a.  A coupling cable  73  is attached to the other end of gate lever  18  by one end and partially wrapped around a cable cam  73   a.  Coupling cable  73  then extends perpendicular to sidewall  13   a  and away from it. Coupling cable  73  can be made from stranded steel or stainless steel miniature cable, or a synthetic polyester fiber string such as DACRON® by DuPont. The relative position of coupling cable  73  is such that the axis of it&#39;s extended portion and the pivot axes of gates  15  and  16  are substantially in the same plane when no pressurized air is supplied to the air volume regulator. As a result, when air begins to flow through the regulator, the direction of movement of cable cam  73   a  is substantially linear and parallel to sidewall  13   a  and in the downstream direction. 
   Now referring to  FIG. 4   b,  the counterbalance section includes two tracks  24  that are fixedly attached along the length of sidewalls  12   a  and  13   a.  Guided within tracks  24 , two chassis guide plate  25   a  slide parallel to sidewalls  12   a  and  13   a  and are fixedly attached to a chassis  25 . An opening  25   b  is cut in chassis  25  to allow the free passage of gate lever  18 . An airtight sliding seal  59  is inserted around the perimeter of opening  25   b  in the space between chassis  25  and sidewalls  12   a  and  13   a.  It is fixedly attached to chassis  25  and in sliding contact with sidewalls  12   a  and  13   a.  Chassis  25  has an upstream end panel  25   d  into which two pivot slots  25   e  are cut. A shuttle  28  with two pivot tabs  28   a  is inserted into pivot slots  25   e.  Pivot tabs  28   a  are substantially of the same construction as pivot tabs  15   b  and  16   b.  Their pivot axis is substantially parallel to the axes of gates  15  and  16 . A second cable cam  73   b  and the second end of coupling cable  73  are fixedly attached to shuttle  28  with fasteners  56  such that coupling cable  73  is partially wrapped around cable cam  73   b.  The initial center distance between cable cams  73   a  and  73   b,  and their common diameters is determined experimentally and is proportional to the maximum travel of coupling cam  73   a.  As can be seen, the linking of the gate lever  18  with shuttle  28  by coupling cable  73  is such that when gate lever  18  moves cable cam  73   a  and the first end of coupling cable  73  in a downstream direction, coupling cable  73  pulls on shuttle  28 , rotating it about pivot tabs  28   a  towards sidewall  13   a  and at right angle to the travel of cable cam  73   a.  A counterbalance spring  30  is inserted in a counterbalance spring pivot hole  28   b  on shuttle  28  such that it can freely rotate from a position perpendicular to shuttle  28  to a position substantially parallel to it. 
   A low friction sleeve bearing  33  is fixedly mounted through chassis  25  such that its axis of rotation is parallel to the axis of shuttle  28 . Although some experimental fine-tuning is required to determine the precise location of the axis of sleeve bearing  33  on chassis  25 , its axis of rotation is in close proximity to the axis of hole  28   b.  An actuator shaft  32  with a spring arm  31  fixedly mounted to its distal end is inserted into sleeve bearing  33  to freely rotate. A threaded pivot bolt  34  is fixedly attached using a locking nut  35  to the distal end of spring arm  31 . An extension nut  36  is screwed onto the end of pivot bolt  34  such that it can freely rotate approximately 45 degrees once in place. A hole is drilled in the distal end of extension nut  36  perpendicular to its axis into which a threaded eye bolt  37  is inserted with a sliding fit. An adjusting nut  38  is added to retain eye bolt  37  within the hole in extension nut  36 . With one end of counterbalance spring  30  mounted in hole  28   b  as outlined above, its opposite end is inserted in the eye of eye bolt  37 . 
   An optional actuator  70  of know construction is shown mounted to actuator shaft  32  on the exterior of casing  25 . 
   Experimentation has shown that counterbalance spring  30  requires more initial tension at maximum airflow. This progressive increase can be accomplishes by relocating actuator shaft  32  away from hole  28   b  in a perpendicular direction from shuttle  28 . Referring to  FIG. 3   d,  the extended length of counterbalance spring  30  varies as the distance between extension nut  36  and hole  28   b.  At minimum airflow with spring arm  31  parallel to shuttle  28 , this distance is equal to the distance between extension nut  36  and shaft  32 . Referring to  FIG. 3   e,  at maximum airflow with spring arm  31  perpendicular to shuttle  28 , the distance between extension nut  36  and hole  28   b  increases by the relocation distance between hole  28   b  and shaft  32 . 
   An alternate arrangement is shown in  FIGS. 3   f  and  3   g.  Again, actuator shaft  32  is relocated away from hole  28   b  but parallel to shuttle  28  and away from shuttle pivot tabs  28   a  and the extended length of counterbalance spring  30  varies as the distance between extension nut  36  and hole  28   b.  Referring to  FIG. 3   f,  at minimum airflow with spring arm  31  parallel to shuttle  28 , this distance is equal to distance between extension nut  36  and shaft  32  minus the relocation distance of shaft  32 . Referring to  FIG. 3   g,  at maximum airflow with spring arm  31  perpendicular to shuttle  28 , the spring  30  extends to the distance between extension nut  36  and shaft  32 . 
   Returning to  FIG. 4   b,  a guide angle  39  is fixedly attached to sidewall  12   a  by fasteners  58  and in which two arms  39   a  and  39   b  are formed. Guide angle arms  39   a  and  39   b  extend outwardly from sidewall  12   a  through opening  25   b  such that one arm is on each side of actuator shaft  32 . An eccentric circular cam  40  is fixedly attached to spring arm  31  and actuator shaft  32 . Eccentric cam  40  is made of a wear resistant—low friction material such as brass, nylon or ultra high molecular weight (UHMW) polypropylene, its diameter is equal to the distance between arms  39   a  and  39   b  and it rotates about its eccentric axis in a sliding fit between arms  39   a  and  39   b.  As spring arm  31  is rotated from its maximum setpoint position perpendicular to shuttle  28 , to a position parallel to it, eccentric cam  40  pushes against arm  39   a  and, guided by tracks  24 , slides chassis  25  and all the components that are attached to it in the downstream direction. Conversely, as spring arm  31  is rotated from its minimum setpoint position parallel to shuttle  28  to its maximum position perpendicular to it, eccentric cam  40  pushes against arm  39   b  and, guided by tracks  24 , slides chassis  25  and all the components that are attached to it in the upstream direction. 
   A flywheel drive bow  43  is fixedly attached to shuttle  28  with fasteners  57 . For clarity, drive bow  43  is shown attached to shuttle  28  at its distal end but could be attached at any convenient location along it. A tensioning spring  44  and a drive cable  45  are strung between the ends of drive bow  43  at two cable hooks  43   a.  Drive cable  45  is kept taut by tensioning spring  44 . Drive cable  45  is flexible and is rolled one or more times around a drive bushing  46  as a string around a toy top. Drive bushing  46  is part of a dampener flywheel assembly  47 . Cable hooks  43   a  are positioned at equal distances from the axis of pivot tabs  28   a.  The position of the axis of flywheel assembly  47  is such that drive cable  45  is substantially tangent to drive bushing  46  as drive bow  43  rotates about pivot tabs  28   a.    
     FIG. 4   f  shows an exploded view and  FIG. 4   g  shows a cross-sectional view of dampener flywheel assembly  47 . It includes a flywheel shaft  55  with a tubular guide portion  55   a  and a flanged portion  55   b.  The following parts are inserted consecutively onto guide portion  55   a:            (a) a flywheel disk  53  such that it rests against flanged portion  55   b  and rotates freely on guide portion  55   a,      (b) drive bushing  46  such that it slides against flywheel disk  53  and rotates freely on tubular guide portion  55   a.  Drive bushing  46  is made of a wear resistant elastomer material having a high coefficient of friction such as urethane or neoprene similar to stripper springs used in tool and die fabrication.   (c) a spring alignment shoulder washer  51  such that it rotates freely on guide portion  55   a.  Shoulder washer  51  is made of a wear resistant—low friction material such as brass, nylon or ultra high molecular weight (UHMW) polypropylene.   (d) a compression spring  52 .   (e) a second spring alignment shoulder washer  51  such that it rotates freely on guide portion  55   a.          
   A flywheel pivot pins  54  having a pointed conical ends  54   a  and a blunt end  54   b  is inserted into flanged portion  55   b  by its blunt end  54   b.  A retaining ring  50  is inserted with a friction fit onto the conical end  54   a  of a second pivot pin  54 . The blunt extremity  54   b  of the second pivot pin  54  is slidably inserted into guide portion  55   a.  Retaining ring  50  comes to rest against shoulder washer  51 , pushing it against spring  52  to compress it. In turn, drive bushing  46  is pushed against flywheel disk  53 . The pressure applied by spring  52  is such that as drive bushing  46  is rotated, it will tend to rotate flywheel disk  53  due to the friction between them. As shown in the typical cross-section through dampener flywheel assembly  47  in  FIG. 4   g,  it rotates freely on conical ends  54   a  of the two pivot pins  54  retained between two conical cup bearings  48 . A conical recess  48   a  is formed in each conical cup bearings  48  to receive pivot pins  54 . The angle of conical recess  48   a  is greater than the angle of pivot pin conical ends  54   a.  For convenience, the exterior of conical cup bearings  48  is threaded and they are fixedly attached to chassis  25  with a nut  49 . The distance between the two conical cup bearings  48  is such that spring  52  is compressed, pushing flywheel pivot pins  54  into conical cup bearings  48  and eliminating all play. 
   Referring to  FIG. 4   e,  an isometric view of an airflow setting quadrant assembly  60  is shown (not shown in  FIG. 4   b ). It is situated on the exterior of chassis  25  and inserted onto actuator shaft  32 . A flow indicator arm  61  with a pointed end  61   a  is fixedly attached to actuator shaft  32 . Its angle of rotation is set by two “U” shaped adjustable limit stops  62  positioned on a slotted quadrant  63 . Limit stops  62  are locked in place using a carriage bolt  64  and a wing nut  65 . Slotted quadrant  63  is attached to chassis  25  by two fasteners  66 . To provide accurate positioning of limit stops  62 , a leg  62   a  is added to limit stop  62  and 
     FIG. 4   h  shows the air volume regulator viewed from below with optional actuator  70  and airflow setting quadrant assembly  60  installed.  FIG. 4   i  shows the air volume regulator viewed from below with only quadrant assembly  60  installed. In  FIGS. 4   h  and  4   i , a removable shroud  26  with an airtight seal at its perimeter  69  (shown in  FIG. 4   b ) is attached over access opening  25   c  in chassis  25  (shown in  FIG. 4   b ) using fasteners  27 . 
   OPERATION—PREFERRED EMBODIMENT—FIGS.  5   a  to  5   e    
     FIG. 5   a  shows a horizontal section through the air volume regulator as the air flows through it and set for its maximum flow rate. A pressure graph is laid out above it to illustrate the variations in pressures at various points as the air travels through the regulator where:
         Point A is taken at the entrance to transition section  11 ,   Point B is taken at the entrance to the flow constricting section  12 ,   Point C is taken at the end of the curved portion of gates  15  and  16 ,   Point D is taken at the narrowest portion or throat,   Point E is taken at the exit of flow constricting section  12 .   Point F is taken within outlet section  13 .       
   Entering the regulator at inlet  10 , the airflow passes from Point A to Point B, expanding in transition section  11  of included angle  11   a.  Transition section  11  advantageously increases the static pressure by converting a portion of the dynamic pressure to static pressure. The efficiency of the conversion is 67% or better if included angle  11   a  of transition section  11  is less than 45 degrees (as per the ASHREA—1989 Book of Fundamentals, page 32.30, table 4-5). The higher static pressure at Point B will be advantageously used to control the airflow as it enters flow constricting section  12 . Two plenums  12   c  and  12   d  formed between gate  15  and sidewall  12   a,  and gate  16  and sidewall  12   b  respectively are pressurized to the same pressure as at Point B. 
   At Point B, airstream  17  begins to constrict and divide into two as it impinges on baffle  19  and rounded upstream ends  15   a  and  16   a  of gates  15  and  16  respectively. Gates  15  and  16  and the apex of baffle  19  form two venturi: two converging passageways, which at their narrowest, are the throat of the venturi. Between Points B and C, the following occurs:
         (a) the air velocity and its associated dynamic pressure increase,   (b) the total pressure remains substantially constant and   (c) the static pressure decreases by substantially the same amount that the dynamic pressure increased.       

   Moving from Point B to Point D, the air velocity increases to the point where its associated dynamic pressure exceeds the total pressure. This generates a negative static pressure or light vacuum in the space between gates  15  and  16  and baffle  19 . The maximum static pressure differential across gates  15  and  16  is at Point D and is equal to the static pressure in plenums  12   c  and  12   d  minus the static pressure at Point D. This is shown on the graph on  FIG. 5   a  as being equal to ΔP. This pressure differential generates a force, which tends to urge gates  15  and  16  towards baffle  19  and restrict the airflow. Baffle  19  helps to lengthen the high velocity segment of the passageways and thus the zone of light vacuum. Ogival shaped baffle air guide  19   b  reduces the noise generated by the expanding air between Point D and Point E as it passes through openings  19   a  and converts a portion of the dynamic pressure back to static pressure as the air expands (“static regain”). 
   Although the static pressure differential across gates  15  and  16  drops between Point D and Point E, it still contributes to urging them towards baffle  19  and restrict the airflow. The relatively large surfaces of gates  15  and  16  over which the above defined pressure differentials are applied generate sufficient forces to bring the airflow under control at pressures of 25 pascals (0.1″ w.g.) or less throughout the full airflow range of the air volume regulator. 
   As the static pressure at inlet  10  increases beyond the 25 pascals (0.1″ w.g.) threshold, the flow restricting force generated by the static pressure differential across gates  15  and  16  exceeds the equilibrating force of counterbalance spring  30 . As cable cam  73   a  starts to move in the downstream direction, coupling cable  73  pulls shuttle  28  rotating it about its pivot tabs  28   a.  The force required to pull shuttle  28  and extend counterbalance spring  30  is at first very low due to the high angle of incidence of coupling cable  73  to the shuttle  28  (almost perpendicular—see  FIG. 5   a ). As the pressure increases, the angle of incidence of coupling cable  73  decreases and the distance traveled by gate lever  18  reduces for an equal incremental pressure rise. At the maximum pressure, the angle of incidence of coupling cable  73  is less than 45° as is shown in  FIG. 5   b.  The net effect is that for a linear increase in the pressure upstream, there is an exponential reduction in the travel of cable cam  73   a.    
   Referring back to  FIG. 1 , this is the desired exponential characteristic for the reduction of the throat of the venturi. Since the area of the air passageway is equal to the height of gates  15  and  16  times the throat width and that the gate height is fixed then the exponential equation for the passageway width is: 
   
     
       
         
           
             Width 
           
           = 
           
             
               constant 
             
             
               
                 static  pressure  drop 
               
             
           
         
       
     
   
   Moving to  FIG. 5   c,  it shows a horizontal section through the air volume regulator at a reduce flow rate and minimum pressure. To reduce the flow rate, three adjustments are made:
         (a) The initial width of the air passageway at points D (the venturi throat) is reduced proportional to the desired airflow reduction.   (b) The spring rate of counterbalance spring  30  is reduced proportional to the desired airflow reduction.   (c) The initial spring tension of counterbalance spring  30  is reduced proportional to the desired airflow reduction.       

   The initial reduction in the venturi throat defines the new starting point of the airflow control and the proportional reduction of the spring rate (softer spring) combined with the reduction of the spring initial tension, maintains the reduced airflow rate substantially constant as the pressure differential varies. 
   All the adjustments are made simultaneously by the rotation of actuator shaft  32 . To reduce the initial throat width of the venturi proportional to the desired flow rate reduction, gates  15  and  16  are made to initiate their flow constricting function proportionally closer to baffle  19 . 
   As actuator shaft  32  rotates:
         (a) eccentric cam  40  rotates and pushes against guide angle arm  39   a,      (b) since guide angle  39  is fixedly mounted the sidewall  12   a,  circular cam  40  pushes chassis  25  with all the components mounted to it (and more specifically cable cam  73   b ) in the downstream direction guided by tracks  24 ,   (c) since coupling cable  73  links gate lever  18  to shuttle  28 , gate lever  18  also move in the downstream direction allowing the pressure differential across gates  15  and  16  to push them towards baffle  19 .       

   As previously outlined in the discussion of prior art, the spring rate of counterbalance spring  30  must vary proportionally to the changes in air volume flow rate for the air volume flow rate to remain substantially constant as the pressure at the inlet to the regulator varies. 
   In referring to  FIGS. 3   a,    3   b  and  3   c,  the spring rate is defined as the “force-displacement” characteristic of a spring. A close approximation of the required variation of the spring rate is advantageously achieved by a simple mechanism that consists of:
         (a) an output shuttle having a fixed linear trajectory to output the desired “force-displacement” characteristic,   (b) a spring pivotably attached at on end to the shuttle and having a predetermined initial deflection and spring rate or “force-displacement” characteristic,   (c) varying the angle of incidence of the spring to the shuttle trajectory.       

   While maintaining the shuttle&#39;s load displacement distance “D” constant in  FIGS. 3   a,    3   b  and  3   c  and referring more specifically to  FIG. 3   a,  the in-line or parallel position is first analyzed:
         the spring reaction force “R” equals the output force “F” (R i =F i  and R f =F f ),   the spring deflection “d” equals the load displacement “D” and   the shuttle output “force-displacement characteristic” K fd  equals the spring rate of the installed spring.       

   Moving to  FIG. 3   b,  it shows the spring rotated to an angle of A degrees. The shuttle&#39;s output force and output “force-displacement characteristic” vary as follows:
         the output force is given by:
           R x =spring load×cosine A=R f ×cosine A   
           the spring load is equal to:
           R f =spring constant×d   
           the spring deflection d is:
           d=D×cosine A   
               

   Combining the three previous equations,
         the output force is equal to:       

   
     
       
         
           
             
               
                 
                   R 
                   x 
                 
                 = 
                 
                   
                     [ 
                     
                       
                         spring  constant 
                       
                       × 
                       
                         ( 
                         
                           D 
                           × 
                           
                             cosine 
                           
                           ⁢ 
                           
                               
                           
                           ⁢ 
                           A 
                         
                         ) 
                       
                     
                     ] 
                   
                   × 
                   
                     cosine 
                   
                   ⁢ 
                   
                       
                   
                   ⁢ 
                   A 
                 
               
             
           
           
             
               
                 = 
                 
                   
                     spring  constant 
                   
                   × 
                   D 
                   × 
                   
                     
                       ( 
                       
                         
                           cosine 
                         
                         ⁢ 
                         
                             
                         
                         ⁢ 
                         A 
                       
                       ) 
                     
                     2 
                   
                 
               
             
           
         
       
     
       
       
         
           the output “force-displacement characteristic” K fd  is given by: 
         
       
     
  
   
     
       
         
           
             
               
                 
                   K 
                   fd 
                 
                 = 
                 
                   
                     
                       R 
                       x 
                     
                     D 
                   
                   = 
                   
                     
                       
                         
                           
                             spring  constant 
                           
                           × 
                           D 
                           × 
                           
                             (cosine 
                           
                           ⁢ 
                           
                               
                           
                           ⁢ 
                           A 
                         
                         ) 
                       
                       2 
                     
                     D 
                   
                 
               
             
           
           
             
               
                 = 
                 
                   
                     spring  constant 
                   
                   × 
                   
                     (cosine 
                   
                   ⁢ 
                   
                       
                   
                   ⁢ 
                   A 
                   ⁢ 
                   
                     
                       ) 
                     
                     2 
                   
                 
               
             
           
         
       
     
   
   Thus for a given spring rate, the shuttle output “force-displacement characteristic” K fd  varies as the square of the cosine of the spring angle.  FIG. 2  shows a graphical representation of the above mechanism: the output “force-displacement characteristic” variation versus the angle of incidence of the spring. As a reference, the desired true linear or ideal variation of the output “force-displacement characteristic” is presented as a dashed line. Although the function (cosine A) 2  is not a linear function, in practice, this mechanism adequately simulates the desired variation of the output&#39;s “force-displacement characteristic”. The discrepancy between the ideal response and the actual variation is easily compensated for by adjusting scaled decal  67  to show the actual flow rate as a function of the counterbalance spring angle. 
   Returning now to  FIG. 5   c,  in rotating actuator shaft  32  in a clockwise direction, spring arm  31  rotates counterbalance spring  30  around hole  28   b  and since the movement of shuttle  28  is limited by pivot tabs  28   a,  the “force-displacement characteristic” as seen by shuttle  28  varies in relation to the angle of rotation of actuator shaft  32 . The tension of counterbalance spring  30  is also reduced as shown in  FIG. 3   d  or  3   e  or a combination thereof.  FIG. 5   d  shows the regulator under reduced flow and maximum pressure. 
   As outlined in the preferred embodiment, the axis of actuator shaft  32  is positioned as close as is practical to the counterbalance spring pivot axis at hole  28   b.  An advantage is sought from this proximal positioning: if counterbalance spring  30  and spring arm  31  aligned, counterbalance spring  30  will not tend to rotate spring arm  31  about actuator shaft  32  irregardless of the angular position of counterbalance spring  30  as it rotates about pivot hole  28   b.  As is shown in  FIGS. 5   c,    5   d  and  5   e,  some misalignment does occur between the counterbalance spring pivot axis at pivot hole  28   b  because its relative position to actuator shaft  32  varies as the pressure conditions change at inlet  10  of the regulator. The angular misalignment of counterbalance spring  30  and spring arm  31  is limited to around 10° by making spring arm  31  sufficiently long to achieve this limitation. This limits the torque generated by counterbalance spring  30  on actuator shaft  32 . 
     FIG. 5   e  shows counterbalance spring  30  fully rotated by spring arm  31  to the minimum airflow position where counterbalance spring  30  lies on a imaginary line between shuttle pivot tabs  28   a  and pivot hole  28   b  (as per the condition in  FIG. 3   c ); counterbalance spring  30  now generates no retaining force. It is to be noted that the angle of rotation of spring arm  31  has exceeded 90 degrees because of the displacement of shuttle  28 . The two gates  15  and  16 , under the action of the pressure differential across them, close against baffle  19  covering openings  19   a  and shutting off the airflow through the regulator. To achieve this, the rotation of eccentric circular cam  40  must slide chassis  25  a minimum distance in the downstream direction to allow cable cam  73   a  to swing freely until gates  15  and  16  close against V baffle  19 : shuttle  28  swings towards sidewall opening  13   c  and coupling cable  73  is no longer under tension. The required travel of chassis  25  generated by the rotation of eccentric cam  40  is determined experimentally. It is proportional to the travel of cable cam  73   a.  As an example, for a 152 mm (6″) diameter inlet  10 , the travel of cable cam  73   a  is 54 mm (2.125″) and the travel of chassis  25  is substantially equal to 13 mm (½″) or approximately one quarter the travel of cable cam  73   a.    
   Referring now to  FIG. 4   e,  flow indicator arm  61  of quadrant assembly  60  limits or fixes the angular displacement of actuator shaft  32  to set the desired airflow rate(s). When the airflow set-point is variable, actuator  70  (shown in  FIG. 4   h ) positions flow indicator arm  61  between two limit stops  62 : one for the maximum airflow, one for minimum airflow. When the airflow set-point is fixed, two limit stops  62  are pushed tight against both sides of indicator arm  61  to lock it in place. In conjunction with scaled decal  67  glued to chassis  25 , indicator arm  61  permits a direct reading of the air volume being delivered through the airflow regulator in operation. 
   Referring now to  FIGS. 4   f  and  4   g,  dampener flywheel assembly  47  is proposed to control the air volume regulator propensity to flutter, oscillate or pulsate under unstable airflow conditions at its inlet. As taught in prior art, a flywheel can be used to change the natural frequency of an oscillating mechanism. This in itself does not dampen the harmonic oscillation, reduce or stop it since no energy is dissipated; only its natural frequency is changed. Thus there exits a pressure pulse frequency at which the flywheel is of no use. 
   My proposed dampener flywheel assembly  47  is built as a “limited torque” drive. It adds dampening by permitting slippage under moderate to high accelerations or decelerations of the flywheel assembly  47  in the frequency range in which a flywheel alone is ineffective. Under normal operating conditions, flywheel assembly  47  rotates very slowly with substantially no friction as the air volume regulator reacts to slow changes in the pressure at its inlet. 
   If a pressure pulse attains the air volume regulator, the “limited torque” drive of dampener flywheel assembly  47  reacts to:
         a) dissipate a portion of the energy as drive bushing  46  slips on flywheel disk  53 , thus limiting the amount of energy which can be stored in flywheel assembly  47 . This loss of energy dampens the pulsation,   b) desynchronize flywheel assembly  47  from the air volume regulator making them out of phase, i.e. the inertia of flywheel assembly  47  will cause flywheel disk  53  to rotate in a clockwise direction and, because of the slippage, the air volume regulator can be rotating drive bushing  46  in a counterclockwise direction.       

   In practice, the mass and diameter of flywheel disk  53  is adjusted to reduce the natural frequency of the air volume regulator to less than ½ cycle per second. The maximum torque that can be applied to flywheel assembly  47  is limited by the force of compression spring  52  pushing drive bushing  46  against it and the friction coefficient between them. The load applied by compressing spring  52  is adjusted by increasing or reducing its deflection. This is achieved by moving internal tooth retaining rings  50  along flywheel pivot pins  54 . The friction coefficient is fixed by the choice of materials to fabricate flywheel disk  53  (usually steel) and drive bushing  46 . For drive bushing  46 , the preferred material choice is an elastomer plastic such as neoprene or urethane that have the required high friction coefficient and a good wear resistance. Compression spring  52  has a dual function: the first one outlined above is to push drive bushing  46  and flywheel disk  53  together to increase friction between them; the second is to push flywheel pivot pins  54  into conical cup bearings  48 , eliminating the need for adjustment between them. 
   DETAILLED DESCRIPTION—ADDITIONAL EMBODIMENTS—FIGS.  6   a,    6   b,    7   a,    7   b,    8   a,    8   b  and  9   
   In  FIGS. 6   a  and  6   b,  the elements of the flow control section are rearranged. Baffle  19  is separated into two along its axis of symmetry that runs through its apex. The first baffle half including baffle arm  19   d  and its associated half of curved baffle air guide  19   b  is then fixedly attached to sidewalls  12   b.  The second baffle half including baffle arm  19   e  and its associated half of curved baffle air guide  19   b  is then fixedly attached to sidewalls  12   a.  Gate  15  remains with baffle arm  19   d  and gate  16  remains with baffle arms  19   e.  Baffle arms  19   d  and  19   e  are then reassembled and sealed together at their downstream edges. The defining characteristics of baffle arms  19   d  and  19   e,  baffle openings  19   a  and the curvature of baffle air guides  19   b  remain the same as in the preferred embodiment. Drive pin  20  and its associated slot  21  are replaced. In their place, 4 links  18   a,    18   b,    18   c  and  18   d  are added such that gates  15  and  16  continue to move in unison and in opposite directions. Links  18   a,    18   b,    18   c  and  18   d  are pivotably attached to each other at their ends, to gate lever  18  and to gate brackets  15   d  and  16   d  with pivot screws  68 . Gate lever  18  and link  18   d  are pivotably attached to sidewalls  13   a  and  13   b  respectively with additional pivot screws  68 . The counterbalance section is the same as the preferred embodiment shown in  FIG. 4   b.    
   Now referring to  FIG. 7   a  and  FIG. 7   b,  two alternatives to coupling cable  73  with cable cams  73   a  and  73   b  are shown. In  FIG. 7   a,  a coupling link  71  is shown pivotably attached by pivot screws  68  to gate lever  18  and shuttle  28 . Coupling link  71  functions in a similar fashion to coupling cable  73 . Referring to  FIG. 7   b,  a second alternative is a cam/cam follower combination which gives similar load transmitting characteristics as coupling cable  73 . A follower bearing  23  is attached to the distal end of gate lever  18 . The relative position of follower bearing  23  is such that its axis of rotation and the pivot axes of gates  15  and  16  are substantially in the same plane when no pressurized air is supplied to the air volume regulator. As a result, when air begins to flow through the regulator, the direction of movement of follower bearing  23  is substantially linear, parallel to sidewall  13   a  and in the downstream direction. A concave circular cam  29  is fixedly attached to shuttle  28  with fasteners  56  such that follower bearing  23  is in rolling contact with the concave circular surface of cam  29 . The radius of cam  29  is determined experimentally and is proportional to the maximum travel of follower bearing  23 . As an example, for a 152 mm (6″) diameter inlet  10 , the travel of follower bearing  23  is 54 mm (2.125″) and the radius is equal to 63 mm (2.5″) with a 23 mm-0.905″ diameter follower bearing  23  or substantially equal to 1.2 times the travel of follower bearing  23 . Cam  29  is so oriented that when gate lever  18  and follower bearing  23  move in the downstream direction, follower-bearing  23  rotates cam  29  about pivot tabs  28   a  at right angle to the travel of follower bearing  23 . 
     FIG. 8   a  shows the proposed adjustable counterbalance spring assembly and cable driven flywheel, as taught in the preferred embodiment, advantageously applied to a known air volume regulator design. Such prior art air volume regulator are shown in U.S. Pat. No. 3,942,552 to Logsdon (1976), U.S. Pat. No. 3,939,868 to Logsdon (1976), U.S. Pat. No. 3,425,443 to Smith (1969), U.S. Pat. No. 3,060,960 to Waterfill (1962), 2890,716 to Werder (1959) and my own Patent 4,130,132 (1978). 
   An airframe is formed by 2 impervious end walls  122 , 2 impervious sidewalls  117 , 2 pervious pitched sidewalls  116 , an inlet opening  126  and a crown  112 . Pervious walls  116  can be a perforating sheet, an assembly of rods or a screen material such that they permit the passage of air. An airstream  17  enters the airframe through opening  126  and exits through pervious sidewalls  116 . The flow restricting gates take the form of two impervious flexible curtains  115  mounted to rigid curtain frames  114 . Curtain frames  114  are pivotably attached by pivot arms  114   e  to end walls  122  near crow  112  with pivot pins  113 . Curtains  115  are fixedly secured to curtain frames  114  at their distal upstream edges  114   d  and are fixedly attached to the air frame at crown  112 . 
   Dampener flywheel assembly  47  is positioned upstream in airstream  17  between curtain frames  114  and walls  122 . Dampener flywheel assembly  47  is substantially the same as shown in  FIGS. 5   a  and  5   b  of the preferred embodiment. In  FIG. 8   a,  cup bearings  49  (not shown) are mounted in end walls  122  such that they are centered between both sidewalls  117 . To restrain curtains  115  and they supporting curtain frames  114 , an equalizer cable  118  is fixedly attached to both curtain frames  114  at  114   c.  A shuttle  111  movably connects equalizer cable  118  at its center to a counterbalance spring  30 . A shuttle hook  111   b  having a circular convex surface is required to preclude the premature failure of cable  118  from flexural fatigue as it flexes at its center. A convex surface is also provided on frames  114  at  114   f  to again preclude cable  118  from breaking by flexural fatigue at its pivot points. 
   The opposite end of tension link  111  is pivotably attached to the end-loop of a counterbalance spring  30  by pivot hole  111   a.  A mounting plate  123  is fixedly attached perpendicular to crow  112  and parallel to end wall  122 . A retaining arm  110  is pivotably attached at one end to mounting plate  123  with pivot pin  110   a.  The distal end of retaining arm  110  is pivotably inserted onto the end-loop of counterbalance spring  30  at pivot hole  110   b.  Retaining arm  110  retains tension link  111  so that it moves in a direction substantially perpendicular to crown  112 . 
   A low friction sleeve bearing  33  is fixedly moounted through mounting plate  123  such that its axis of rotation is parallel to crown  112  and intersects the centerline of tension link  111 . Some experimental fine tuning is required in determining the precise distance of the axis of sleeve bearing  33  from crown  112 . Its axis of rotation is above the collinear axes of pivot holes  110   b  and  111   a  when the airflow regulator is not in operation. As with the preferred embodiment, an actuator shaft  32  with a spring arm  31  fixedly mounted to its end is inserted into sleeve  33  to freely rotate. An optional actuator  70  (not shown) of know construction or airflow setting quadrant assembly  60  (as shown in  FIG. 4   e ) or both can be mounted to the opposite end of actuator shaft  32  to position spring arm  31  and set the flow rate. The following parts are attached to spring arm  31  in the same way as with the preferred embodiment: a threaded pivot bolt  34  (not shown), a locking nut  35  (not shown), an extension nut  36 , an threaded eye bolt  37  and an adjusting nut  38 . 
   With one end of counterbalance spring  30  mounted in pivot holes  110   b  and  111   a  as outlined above, its opposite end is inserted in the eye of eye bolt  37 . A scaled air volume decal  67  is fixedly attached to mounting plate  123  to permit a direct reading of the airflow set point. 
   Now referring to  FIGS. 8   a,    8   b  and  8   c,  a novel method of maintaining the synchronized operation of the pair of curtain frames  114  is shown. Using 2 drive cables  124   a  and  124   b,  this method makes curtain frames  114  move in unison and in opposite directions and also drives dampener flywheel assembly  47 . Two drive bows  119  and  120  are pivotably fixed near the upstream ends of curtain frames  114  at  114   a  and  114   b  respectively. A tensioning link  121  is pivotably attached to the distal end of drive box  120  at  120   a.  As shown in  FIG. 8   b,  drive cable  124   a  is strung from drive bow  120  at  120   b,  around drive bushing  46 , to drive box  119  at  119   a.  As shown in  FIG. 8   c,  the second drive cable  124   b  is strung from drive bow  119  at  119   b,  around drive bushing  46 , to tensioning link  121  at  121   a.  For practical reasons, the geometry of drive bows  119  and  120  is such that drive cables  124   a  and  124   b  are made the same length. To complete the assembly, one end of a tensioning spring  125  is hooked to drive bow  120  at  120   c  and its distal end is hooked on tensioning link  121  at  121   b.  This keeps drive cable  124   b  taut. Because drive cable  124   a  is attached to both drive bows, tensioning spring  125  also keeps drive cable  124   a  taut. 
   An alternate to eccentric cam  40  and its associated guide angle  39  is shown in  FIG. 9 . A connecting link  41  is pivotably attached at both ends by two pivot screws  68 , one end to spring arm  31  and the other to a retaining angle  42 . Retaining angle  42  is fixedly attached to sidewall  12   a  by fasteners  58  and extends through opening  25   b  out passed actuator shaft  32 . As spring arm  31  is rotated, connecting link  41  pulls or pushes chassis  25  and all the components that are attached to it, such that the desired relationship between the travel of chassis  25  and the rotation of spring arm  31  is maintained as per the preferred embodiment. 
   OPERATION—ADDITIONAL EMBODIMENTS—FIGS.  6   b,    7   a,    7   b,    8   a,    8   b  and  9   
   Referring to  FIG. 6   b,  an additional advantage is gained from modifying the arrangement of V baffle  19  and gates  15  and  16 . Placing gates  15  and  16  in the center of airstream  17 , the pressure which tends to push them towards baffle  19  is increased by the dynamic pressure of the airstream at their upstream ends  15   a  and  16   a  as airstream  17   d  enters the space between them. Since the airflow is maintained constant through an air volume regulator so is the dynamic pressure and the net pressure increase remains constant as the static pressure at the inlet varies. Thus no adjustments are required in the response characteristics of the linkage and counterbalance spring  30  is selected slightly stiffer. This increase can be substantial at high inlet velocities (at maximum airflow capacity) but is of limited effect at low inlet velocities (at minimum airflow capacity). This makes the regulator minimum static pressure at maximum airflow capacity less than the minimum static pressure at minimum airflow capacity. This situation is the inverse of what is normally seen in airflow regulators. Tests have shown that values of the regulator minimum static pressure at maximum airflow capacity can approach zero. Thus, with these conditions, the sum of the pressure regain generated by transition section  11  and the dynamic pressure entering the space between gates  15  and  16  can be sufficient to initiate airflow control of the airflow regulator and maintain a substantially constant flow of air through it. 
   Now referring to  FIG. 7   a  and  FIG. 7   b,  alternatives to coupling cable  73  with cable cams  73   a  are shown. Coupling link  71  and follower bearing  23  with cam  29  generate the same “force-displacement” characteristic between gate lever  18  and shuttle  28  as compared to the use of coupling cable  73  with cable cams  73   a.    
   Referring to  FIGS. 4   f  and  4   g  of the preferred embodiment, the torque driving flywheel assembly  47  is adjusted by increasing or reducing the slippage between drive bushing  46  on flywheel disk  53 . The slippage is controlled by adjusting the deflection of compression spring  52 . An alternate means of limiting the torque delivered to flywheel disk  53  is to fix drive bushing  46  to flywheel disk  53  and adjust the tension of drive cable tensioning spring  44 . This will allow some slippage of drive cables  45  on drive bushing  46 . The net effect will be the same: if the torque delivered by drive cable  45  exceeds a given amount, it will slip and dissipate a portion of the energy. Drive cable  45  can be coated with a wear resistant material such as nylon. Also, drive bushing  46  can be made of a wear resistant material such as nylon, brass or ultra high molecular weight (UHMW) polypropylene. Optionally, internal tooth retaining ring  50 , spring alignment shoulder washer  51  and compression spring  52  could still be used to eliminate the play between pivot pins  54  and conical cup bearings  48 . In this alternate method, no change in performance is seen as compared to dampener flywheel assembly  47  of the preferred embodiment. 
   Referring to  FIGS. 8   a,    8   b  and  8   c,  the application of the adjustable counterbalance spring assembly, as taught in the preferred embodiment, solves a long felt need: a simple means to adjust the spring rate of the counterbalance spring. Numerous attempts have been made over time to create a low cost and efficient adjustable spring rate mechanism as proven by the numerous U.S. Pat. Nos.: 3,942,552 to Logsdon (1976), 3,939,868 Logsdon (1976), 3,425,443 to Smith (1969), 3,060,960 to Waterfill (1962), 2890,716 to Werder (1959) or my own Patent 4,130,132 (1978). 
   The “limited torque” cable driven flywheel also has several major advantages:
         (a) Less moving parts reduces the associated friction, which allows the air volume regulator to initiate and maintain control of the airflow using less pressure. The air volume regulator minimum static pressure is advantageously reduced.   (b) The use of a cable rather than a linkage and pivot screws also reduces the friction by eliminating sliding surfaces inherent to pivot pins or pivot screws.   (c) The “limited torque” characteristic of proposed dampener flywheel assembly  47  adds the required energy dissipation to the dampening the airflow induced oscillations of the air volume regulator without hindering its flow tracking characteristic.       

   CONCLUSION, RAMIFICATIONS AND SCOPE OF THE INVENTION 
   Accordingly, the reader will see that the “airflow powered” air volume regulator of this invention reliably regulates the flow of air passing through it at pressures of 25 pascals (0.1″ w.g.) or less and this while respecting the industry standard variation of +/−5%. In addition, it will do this over a wide airflow range without having to change or manually adjust the installed counterbalance spring. With the use of the optional actuator, the airflow can be shut-off (zero flow) if desired. When pneumatic actuators are selected, they will not “slide” the airflow set point as the pressure in the ductwork system varies. With the control mechanism situated outside of the airstream, it is not affected by the accumulation of airborne particles on the moving parts. The propensity to pulsate is controlled with the use of the “limited torque” drive flywheel. Furthermore, once installed in the ductwork system, its airflow set point is fully adjustable over the airflow range without having to open access panels or the use of any tools. Adding an optional actuator is simple and easily done without affecting the calibration of the unit or having to open access panels. The minimum and maximum airflow rates are easily adjustable at any time, again without affecting the calibration of the unit, having to open access panels or the use of any tools. 
   While may above description contains many specificities, these should not be construed as limitations on the scope of the invention, but rather as an exemplification of a preferred and alternate embodiments thereof. Other variations are possible. 
   Accordingly, the scope of this invention should be determined not by the embodiments illustrated, but by the appended claims and their legal equivalents.