Abstract:
A bi-fuel and dual-fuel engine variable pressure fuel system is presented facilitating individual or simultaneous use of liquid and gaseous fuels including natural gas, hydrogen and gasoline, through employment of a variable output pressure gaseous fuel regulator incorporating an attached hydraulic amplifying structure communicating with a relatively low pressure fluid servo circuit that may in turn communicate with a variable pressure automotive liquid fuel system to facilitate relatively high pressure gaseous fuel injection.

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     The instant application is a continuation-in-part of U.S. patent application Ser. No. 13/374,810, filed Jan. 14, 2012. The disclosures of the application are incorporated herein by reference. 
    
    
     BACKGROUND OF THE INVENTION 
     A variable pressure gas rail injector pressure means is described by Willi in U.S. Pat. No. 5,771,857, as applied to direct injection, glow ignited natural gas engines. Here variable gas rail pressure is generated by an electronically modulated diesel injection pump that applies high pressure diesel fuel to the control side of a dome loaded regulator to produce correspondingly high, unamplified variable gas injection near TDC to optimize variable pressure direct injection. 
     Laing and Prichard in Canadian patent CA1203132 describe a duel fuel diesel engine, utilizing variably reduced hydraulic pressure in a servo circuit bled from the diesel injection pump and controlled by a centrifugal governor which variably pressurizes the control or load chamber of a gas regulator with diesel fuel in typical dome load fashion, to provide variable gas fuel pressure to a diesel engine air intake. 
     Bickley in U.S. Pat. No. 7,178,335 describes a spool valve hydraulic pressure regulator with variable output pressure controlled by a hydraulic load chamber augmented by an internal load spring whose compressive force is varied by an abutting moveable piston adjustable by means of a separate hydraulic actuating chamber contiguous with the end of the piston opposite the spring. 
     McMahon and O&#39;Halloran in U.S. Pat. No. 7,922,833 describe an invention utilizing a hydraulic cylinder attached to a gas regulator that contains a piston displaceable against a point on a flexible regulator pressure sensing diaphragm for the purpose of varying the tension within the diaphragm in order to vary the pressure of the outflowing gas flowing into a deburring thermal energy machine (TEM.) Variable regulator outflow pressure is here determined by varying the tension of the flexible diaphragm, as opposed to the present invention, where variable hydraulic servo pressure acting through an amplifying piston-pushrod structure upon a regulator pressure sensing piston is the regulator load controlling element as opposed to variable tension within a regulator sensing diaphragm. 
     Multiple variable pressure regulator control means including pneumatic, hydraulic, mechanical, electric and electro-hydraulic are cited in the ECU controlled variable gas pressure system of King in U.S. Pat. No. 5,367,999. A detailed description is provided describing a variable pressure pneumatic regulator actuator embodiment in this specification, but only general reference is made to other variable hydraulic pressure regulator biasing means in the claims, with no details provided in the specification. 
     Douville, Noble, Baker, Tran and Touchette describe a dual fuel diesel direct injection system in U.S. Pat. No. 6,298,833 having one injector that injects both a gaseous main charge and a diesel fuel pilot ignition charge into the engine cylinder, and where a dome loaded regulator directly senses diesel pilot injector fuel pressure, and regulates the main gaseous fuel charge at an equal or slightly lower output pressure, to maintain a positive seal between the gas and liquid fuels within the injector. 
     Post and Brook in Pub. No. US2006/0213488 A1 describe a variable pressure direct gas injection system that includes a hydraulic dome loaded regulator that contains a spring biased flow control valve where the hydraulic load fluid acts against the bias spring to vary gas injector fuel pressure (in a manner similar to McMahon and O&#39;Halloran.) The hydraulic load fluid may consist of diesel pilot fuel and here is always approximately equal to or higher than the regulator outflow gas pressure to avoid gas leakage into the diesel load control fluid. 
     Ancimer, Batenburg and Thompson in U.S. Pat. No. 7,463,967 present a variable pressure, direct supersonic gas injection control system utilizing a single injector for both the diesel pilot and the main gaseous fuels. This also includes a dome loaded regulator that maintains almost equal pressure within the gas and the liquid portions of the injector to insure an effective seal between the two fluids. 
     Palma in U.S. Pat. No. 6,626,150 and Dokas, Pyle and Yu in U.S. Pat. No. 7,624,720 describe electromagnetically controlled gasoline type regulators. 
     Hashemi in U.S. Pat. No. 7,140,354 reveals a means for depressurizing a gaseous fuel injector supply rail with a pump that pumps excess gas from the fuel rail back upstream into either the gas supply tank or to a point upstream of one of the pressure reducing regulators that feed the fuel rail. This pumping means is controlled by an ECU for the purpose of maintaining rail pressures compatible with the operating characteristics of gaseous fuel injectors. 
     The present invention is differentiated from prior art by its&#39; ability to safely utilize low pressure, volatile spark ignitable fuels as a hydraulic regulator servo pressure fluid to produce an amplified, high pressure fuel supply from a gas regulator. The present servo amplifying means differs from conventional dome loaded regulators in that the pressurized servo fluid is mechanically isolated and amplified by the piston-pushrod structure, which moves to block orifices in the present hydraulic amp communicating with the vehicle fuel system in response to a high pressure leak from the gas regulator. 
     The presently described electronic cylinder cutoff and fuel pressure sensing throttle control means obviate the need for the pump dependent, high pressure fuel rail de-pressurizing means described in Hashemi. 
     The present invention is applicable to bi-fuel and dual-fuel internal combustion engines that utilize gaseous and liquid fuels either simultaneously or individually. It is aimed at dealing with the limited response characteristics of high pressure solenoid type gaseous fuel injectors when activated by present 12 volt petrol (gasoline) engine control units (ECU&#39;s,) where the injectors are synchronized to the speed, or RPM of the engine. To compensate for the larger volumes of gaseous fuel required to deliver the equivalent energy of gasoline, gaseous (gas) injectors operate under higher pressures, with larger, heavier moving valve components as compared to petrol type injectors. This can result in minimum open/close cycle periods twice as long as those of their petrol counterparts. At low speed idle power with a static, high fuel rail pressure necessary for maximum power, the gas injector can fail to fully open in response to short ECU commanded voltage pulse widths. Minimum open cycle periods for solenoid gas injectors are typically around 4 milliseconds. At idle with a static gas fuel rail pressure that can meet the engines full operating power range, the ECU may command an injector open pulse width far less than 4 milliseconds. The injector may thus fail to respond fast enough to these short open signals, resulting in inconsistent fuel delivery, roughness and excessive emissions. 
     Where fuel injectors are typically synchronized by the ECU to cycle with engine RPM, low engine speeds allow more time for the injector to more accurately meter fuel. By lowering fuel supply pressure at idle speeds the injector can remain open longer, allowing more accurate response to the ECU. However when fuel demand increases with speed and the available injector cycle time decreases, a variably higher pressure fuel rail supply then becomes necessary to avoid fuel starvation. 
     SUMMARY OF THE INVENTION 
     The operating limitations of solenoid actuated gas injector valves are overcome here by proportionately raising and lowering fuel injector rail pressure with engine speed and load. Longer “open” voltage pulse width commands at low speeds are made possible with lower fuel supply pressures, allowing the injector to deliver small gas quantities per cycle with greater accuracy, while high fuel rail supply pressures are available at maximum speed and load. Controlling fuel flow through statically open gaseous injectors is also made possible through a precisely controlled variable pressure injector rail supply responsive to engine speed and load. 
     The present invention either eliminates entirely, or augments the typical output pressure controlling load spring that acts against the pressure sensing element within a gas regulator (usually consisting of a piston or flexible diaphragm attached to a flow control valve) by utilizing an attachable variable hydraulic pressure amplifying actuator henceforth referred to as a “hydraulic amp.” The simplified regulator represented here is of the 20-1000 compressed natural gas (CNG) type presently made by Tescom, of Elk River, Minn., with a 50-89 psi load spring depicted in  FIGS. 3-3   b.    
     Within the hydraulic amp of the present invention is a hydraulic pressure sensing piston and contiguous pushrod structure referred to henceforth as a “piston-pushrod.” This spool-like structure has a relatively large pressure sensing piston crown surface at the “piston” end, and a smaller surface at the opposite “pushrod” end that either directly abuts the gas regulator&#39;s pressure sensing element on the seat normally acted upon by the load spring, or alternatively, may abut a load spring interspersed between the pushrod and the sensing element. This piston-pushrod structure reciprocates within the hydraulic amp in response to pressure exerted on its&#39; piston crown by a variable pressure hydraulic servo circuit, which in the illustrated embodiment is comprised of a communicating gasoline or diesel liquid fuel supply system having an electric or engine driven fuel or “lift” pump. Other possible sources of servo pressure may be derived from vehicle fluid systems such as a windshield washer system, the engine cooling system and an air pump that supplies braking force or suspension adjustment pressure. One or more variably restrictive flow controlling devices within the servo pressure circuit may include a variable flow control valve and a fuel pressure regulator, to restrict flow and create variable backpressure sensed within the communicating hydraulic amp. An engine control unit (ECU) or separate computers may control servo fluid pressure by sending variable voltages to the fuel pump and variable flow servo control devices in response to engine fuel demand parameters related to speed and load such as manifold absolute pressure (MAP), RPM. Variable backpressure in the present embodiment caused by a fuel pump working against the servo flow control devices results here in a variable servo pressure of approximately 5 to 40 psig. Servo pressure within the hydraulic amp acting upon the piston-pushrod and transmits amplified control pressure directly to the gas regulator pressure sensing piston as shown in  FIGS. 1 and 2 . In an alternate embodiment shown in  FIG. 3  through  FIG. 3 b   , this amplified control pressure may act through the piston pushrod upon a typical load spring that in turn exerts a variable, elastic force upon the sensing piston. In typical fashion as depicted in the accompanying  FIGS. 1 through 3 , a flow control valve within the gas regulator attached to lower surface of the sensing piston variably reciprocates upon a closeable orifice positioned between the regulator&#39;s gas inflow and outflow conduits and meters outflowing gas pressure in response a control or “load” pressure acting on top side of the sensing piston, which in the present invention is ultimately determined by fluid servo pressure amplified and transmitted by the contacting piston-pushrod. Gas regulator outflow pressure here is multiplied over that of the servo fluid pressure by a factor determined by the difference in diameters between the relatively large hydraulic pressure sensing crown of the piston-pushrod within the hydraulic amp, and the smaller abutting gas pressure sensing piston within the regulator. 
     Alternate “Two Stage” Embodiment 
     To facilitate fluctuating servo fluid pressure from sources such as engine driven and electric coolant or fuel pumps that may produce insufficient steady low servo pressure at idle, an alternate embodiment is presented employing a conventional load spring positioned between the piston pushrod and the regulator pressure sensing element, activated by a moveable, surface area augmenting, servo pressure sensitive collar interfacing with crown of the piston pushrod. The collar may be in the shape of a thick flat washer that surrounds the circumference of the piston crown. The collar is fluidly sealed against the inner circumference of the amp housing and the circumference of the crown, and has restricted travel therewith. To move the load spring from an inert position to one producing low range control pressure against full motion of the regulator piston, this collar provides an augmenting force to assist the piston-pushrod in activating the load spring at the lower part of the engines fueling map. As depicted in the vertically disposed  FIG. 3  cross sections, upon exposure to initial low servo pressure, the collar surface increases the effective servo pressure amplifying crown surface area of the piston pushrod providing additional force to induce limited downward motion of the piston pushrod through contact with a flange extending outward from the lower edge of the piston crown, in order to move the load spring into its low active range, initiating “stage one.” During “stage one” of this embodiment, the piston-pushrod crown and the collar acting together apply an amplified servo pressure to hold a limited stationary downward position of the piston-pushrod in order to maintain the load springs&#39; 50 lb. minimum control pressure against the regulator piston throughout its&#39; full travel range. At this point, further motion of the collar and piston-pushrod is blocked by contact of the lower collar surface with the bottom inner wall of the hydraulic amp housing. Further downward movement of the piston-pushrod within the stationary collar ceases until rising servo pressure amplified solely by the smaller area of the piston crown exceeds the 50 lb. counterforce of the activated load spring. The load spring will thus maintain a steady 50 psi gas regulator output pressure until engine fuel demand increases above idle to where rising servo pressure is consistently stable. As servo pressure, amplified solely by the piston-pushrod crown (minus the added force of the now stationary collar) rises to a higher point exceeding the 50 lb. counterforce of the spring, (a pressure margin defined by the difference in surface areas between the piston crown and the collar,) “stage  2 ” activates, urging the piston-pushrod further downward within the stationary collar, applying increasing control pressure through the load spring to the sensing piston, thus raising regulator output pressure. 
     Throttle by Wire 
     Where specific air-fuel ratios are required over a wide range of output power, typical air inlet throttle mechanisms may be more responsive to rapid power change commands than variable output gas regulators, due to the compressibility of the high pressure fuel gas in the piping and fuel rail downstream of the regulator. Air inlet throttle control mechanisms may thus track and respond to transient power demand changes induced by variable pressure delivery gas regulators faster and more accurately than variable pressure delivery regulators can respond to air throttle induced power changes, and may thus produce finer control of air-fuel ratios. In one embodiment of the present invention especially applicable to engines operating within narrow air/fuel ratio limits, a “throttle by wire” system having an intake throttle valve operable in response to variable gas injector rail pressure may be employed to counter compressibility induced rail pressure lag during rapid fuel pressure induced power changes. 
     Precise air/fuel ratio control may therefore be obtained in the present embodiment through employment of a pneumatic or electrically actuated throttle mechanism that responds to operator commanded variable pressure within the gas injector rail. Power output in the present invention thus may be controlled by a “gas pedal” that actuates variable hydraulic flow and pressure controlling components within the present servo circuit (such as a variable flow hydraulic valve), that in turn control gas regulator output and injector rail pressure. Variable gas rail pressure can then operate a pneumatically actuated throttle valve, or be sensed by a throttle controlling ECU that proportionately actuates a motorized throttle valve. 
     Injector/Cylinder Deactivation 
     In multi injector configurations of the present invention, transient fuel rail pressure imbalances resulting from rapid power changes may be countered by employing an ECU injector or injector/cylinder cut off circuit. This circuit may contain a map that defines an injector operating envelope determined by RPM, fuel rail pressure and minimum pulse width. When fuel rail pressures exceed the injector&#39;s minimum pulse width, such as may occur when the operator rapidly lifts off of a fully depressed gas pedal with maximum fuel rail pressure, the ECU may deactivate one or more injectors (and cylinders,) causing the remaining injectors to operate at higher loads with longer pulse widths. When engine load and intake air flow increase, the idle injectors may then be progressively reactivated allowing continuously optimal injector operation and precise air/fuel ratio control. 
     There are numerous superior advantages, as will now be explained. 
     Pressure Droop 
     By eliminating the fixed load spring in a typical gas pressure regulator, the present hydraulic amp embodiment serves to eliminate output pressure drop or “droop” that occurs when control pressure exerted by the load spring on the sensing piston decays as the spring extends as the piston responds while moving the attached flow control valve open. Output pressure thus declines as gas flow demand increases with a typical gas regulator governed by a load control spring. 
     Supply Pressure Effect 
     The force required to open a closed, unbalanced gas regulator flow control valve must exceed the force exerted upon the valve head by the upstream supply tank supply pressure plus that of the force of the out flowing gas pressure upon the pressure sensing piston. When pressure on the inlet (tank) side of the valve falls with fuel consumption, the total force holding the valve closed decreases. Thus, the total force required to open the valve is reduced as the upstream supply tank pressure falls as fuel is consumed. For a conventional gas regulator with a fixed output controlling load spring, the output pressure to the fuel rail will then increase as supply tank pressure decreases. To maintain a constant regulator outflow pressure, the controlling pressure exerted on the load or control side of the regulator sensing piston must be reduced as supply tank pressure decreases. By replacing or augmenting the common regulator load spring with the present servo pressure actuated hydraulic amp, controlled by an ECU having input from an upstream pressure sensor such as a fuel tank quantity gauge, the present invention can maintain consistent outflow pressures independent of falling tank pressure, and eliminate droop associated with a fixed regulator load spring. 
     Expanded Range, Fewer Components, Safety 
     The wide range of controllable gas regulator output pressures (approximately 40 to over 95 psig in this iteration) made possible by the present servo controlled hydraulic amp expands the limited operating bandwidth of solenoid gas injectors. By lowering rail pressure at reduced engine speeds and loads, more accurate metering, lower injector noise and reduced power consumption is attained. Conversely, as RPM increases and the available injector open time per cycle decreases, the present invention increases injector rail pressure with increasing engine speed and fuel demand, increasing fuel flow through injectors that eventually may remain statically open at maximum engine speeds. Employed in a throttle body injection (TBI) configuration, the present variable gas rail pressure invention facilitates the utilization of fewer gas injectors, verses employing a plurality of injectors staged to operate over a wide load and speed range with a constant rail pressure. 
     By replacing or augmenting the regulator load spring in a gas regulator with the present fluid servo pressure controlled amp, when the engine and hydraulic servo pump stop, servo pressure bleeds down and residual gas rail pressure acts unopposed against the regulator&#39;s sensing piston to close the regulator flow control valve. Gas flow to the injectors is then blocked, reducing potential gas leakage through the injectors and the need for a downstream shut off valve typically placed in the conduit running between the regulator and the injector rail. 
     A safety advantage over common dome loaded regulators that have a load or servo control fluid applied directly to the regulator pressure sensing element occurs whereby the servo fluid and communicating vehicle fuel system of the present invention are protected from high pressure gas incursion from a damaged gas regulator sensing element by the present hydraulic amp. 
     Further areas of applicability of the present invention will become apparent from the detailed description provided hereinafter. It should be understood that the detailed description and specific examples, while indicating the preferred embodiment of the invention, are intended for purposes of illustration only and are not intended to limit the scope of the invention. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The present invention will become more fully understood from the detailed description and the accompanying drawings, wherein 
         FIG. 1  is a detailed schematic of the present invention employed in a bi or dual-fuel throttle body injector fuel induction system; 
         FIG. 1 a    is a voice coil actuated iteration of the electromagnetic liquid fuel pressure regulator  56  shown in  FIG. 1 ; 
         FIG. 2  is a detailed cross sectional representation of the hydraulic amp of the present invention. 
         FIG. 3  is a detailed cross sectional representation of the alternate “two stage” iteration of the present invention, deactivated with zero servo pressure; 
         FIG. 3 a    is a detailed cross sectional representation of the “first stage” operational configuration of the alternate iteration of the present invention; and 
         FIG. 3 b    is a detailed cross sectional representation of the “second stage” operational configuration of the alternate iteration of the present invention. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     The following description of the preferred embodiment(s) is merely exemplary in nature and is in no way intended to limit the invention, its application, or uses. 
       FIG. 1  shows a cross sectional representation of the present invention where gaseous and liquid fuels may be injected into the air inlet of an engine at throttle body  24 . Gaseous fuel tank  2  may hold fuels such as natural gas and hydrogen at pressures currently averaging 3600 psig or higher. Liquid tank  4  may hold fuels such as gasoline, alcohol or diesel. Tank  4  supplies variable output fuel pump  7 , which may be electrically or engine driven. Pump  7  supplies petrol injectors  28  within throttle body  24  through fuel line  5  and fuel rail  30 . Throttle body  24  is represented here with two throttle bores  26  having typical shaft mounted throttle valves  29 . Gaseous injectors  22  are mounted opposite liquid injectors  28  in throttle body  24 . Either a gaseous or a liquid fuel can alternately be injected at throttle body  24  through either gas injectors  22  or liquid (“petrol”) injectors  28  in a “bi-fuel” application of the embodiment. In a “dual-fuel” mode, gaseous and liquid fuels may be injected simultaneously by gas injectors  22  and liquid injectors  28 , such as where alcohol or an alcohol solution may be selectively injected with methane or hydrogen in a supercharged application to avoid detonation or knock. 
     Petrol pressure in rail  30  may be controlled by a typical spring loaded bypass regulator communicating with fuel rail  30 , or by the variable pressure electromagnetic bypass regulator shown here at  56 , connected to rail  30  through pipe  57 . Fuel bypassed through regulator  56  returns to tank  4  from output pipe  59  through contiguous fuel return line  54 . Regulator  56  may be electronically controlled to maintain petrol pressure in rail  30  in a typical throttle body or port gasoline fuel injector pressure range of approximately 15 to 75 psig. 
     Fuel tank  2  supplies high pressure gaseous fuel, typically stored at pressures ranging from 200 to over 3600 psig, to piston type gas regulator  10 , through pipe  6 . Pipe  6  contains an electromagnetic shut off valve  8 , and a temperature and pressure sensor  23 . Gas regulator  10  variably reduces storage tank pressure to a range of approximately 40 to 95 psig in the present embodiment to feed gaseous injectors  22  through pipe  18  and rail  20 . Variable output pressure from regulator  10  is produced by attached hydraulic amp assembly  32  that controls regulator pressure output in place of an output governing load spring. Amp  32  is variably pressurized by liquid fuel from pump  7  communicating with internal amp pressure sensing chamber  40  through fuel lines  5  and  9 . Amp  32  may have orifice  48  located near the periphery of sensing chamber  40  to allow communication between it and weep line  49 , and may then be rotated to locate orifice  48  uppermost so as to allow trapped air to rise and pass out of chamber  40  through weep line  49 , and into return line  54  and tank  4 . Tank  4  may then be vented in a typical fashion. Line  49  is of a sufficiently small diameter so as to allow variable servo pressure to be maintained in chamber  40 , while still allowing a small venting flow of fluid into return line  54 . 
     Referencing  FIG. 1  and  FIG. 2 , hydraulic fluid pressure transmitted through lines  5  and  9  through orifice  52  to chamber  40  is sensed through diaphragm  42  (here composed of 1/32 inch thick fluorosilicone rubber.) Diaphragm  42  acts to seal pressurized servo fluid within chamber  40  and transmit variable servo pressure to the piston crown  35  of piston-pushrod  34 . Pressure acting upon crown  35  exerts an amplified force through piston-pushrod  34  upon the load sensing surface  61  of regulator pressure sensing piston  16 . An alternative sealing chamber may consist of an O-ring placed circumferentially between crown  35  of piston-pushrod  34  and the inner adjacent surface of hydraulic amp  32 . Gaseous regulator sensing piston  16  is attached to flow control valve  14 , which is variably closeable upon valve seat  15 . Valve  14  reciprocates upon seat  15  to control the flow of gas from inlet pipe  6  through orifice  21 , into gaseous pressure sensing chamber  19  and out pipe  18 . Variable output pressure from regulator  10  is determined by the reciprocation of valve  14  upon seat  15  in response to the opposing forces acting on attached regulator sensing piston  16  by the outflowing gas pressure in chamber  19  acting on lower piston surface  60 , and the servo pressure in amp chamber  40  acting through structure  34  upon upper piston surface  61 . Upstream supply pressure within valve head chamber  27  may act variably against valve  14  when valve  14  is on or near seat  15 , and can thus add a variable closing force to that of the pressure within chamber  19 . The approximate amplification factor of gas regulator output pressure to liquid servo circuit pressure is determined by the ratio of the surface area of crown  35  of structure  34  to the smaller surface area  60  of regulator piston  16 . Hydraulic servo circuit pressure in the present iteration ranges from approximately 12 to 35 psig, resulting in a gas rail pressure of about 40 to 95 psig. 
     Hydraulic Servo Circuit 
     Referencing  FIG. 1 , variable pressure within amp chamber  40  is determined by a hydraulic servo pressure control circuit comprised here of the following communicating fuel lines and variably restrictive components: Fuel pump  7  which feeds fuel line  5  and branching line  9 ; line  43  branching off of line  9  beneath solenoid valve  11  communicating with variable flow control valve  45 ; line  63  fluidly connecting valve  45  with bypass regulator  47 ; and line  53  fluidly connecting regulator  47  to fuel tank return line  54 . Pressure within chamber  40  can be modulated through variable activation of pump  7  and valve  45 , as well as by electromagnetic regulator  56  communicating with fuel lines  5  and  9  through pipe  57  and liquid fuel injector rail  30 . Bypass regulator  47  is located downstream of valve  45  in order to facilitate stable pressure within the servo circuit at minimal fluid flow, and is set to maintain a minimum servo circuit pressure of 12 psig in this embodiment. Variable flow valve  45  may be comprised of a housing containing an orifice variably closeable by a threaded needle or spool valve reciprocating within a threaded bore, or of a rotating barrel valve, all of which may be actuated by an electronic stepper motor. Valve  45  if of a reciprocating spool or needle configuration, may alternately be actuated by a linear motor. 
     Electromagnetic petrol regulator  56  communicating through pipe  57 , rail  30  and fuel lines  5  and  9 , may control hydraulic fuel pressure to both injectors  28 , and selectively to upstream hydraulic amp chamber  40  through solenoid valve  11 . Regulator at  56  variably reciprocates valve head  72  by an attached armature  62  actuated by a surrounding coil as depicted, or may alternately actuate valve  72  by an attached voice coil moveable within a magnetic field as seen in  FIG. 1 a   . A single variable pressure regulator of sufficient dynamic range at  56  can thus obviate the requirement for separate parallel servo circuit components  45  and  47 . 
     Restating the basic control principal of the invention, variable servo pressure within amp pressure chamber  40  may be regulated by varying the speed and output of pump  7  through electric or engine driven means, and/or by varying the flow capacity of variable valve  45 , and/or by electrically modulating the movement of valve  72  within electromagnetic regulator  56 . Backpressure generated by these components is sensed within amp pressure chamber  40  and amplified by virtue of the relatively large diameter of crown  35  of piston-pushrod structure  34 , versus the smaller diameter of regulator sensing piston  16 . Regulator piston  16 , sensing the amplified force of pressure chamber  40  acting through structure  34 , and the opposing force from regulator output chamber  19 , variably reciprocates connected flow control valve  14  upon orifice seat  15  to deliver a servo controlled variable gas pressure supply to injector rail  20 . 
     Bi-Fuel and Dual-Fuel Modes 
     Referencing  FIG. 1 , in petrol fuel only mode solenoid valve  11  within pipe  9  is closed, allowing hydraulic pressure within amp pressure chamber  40  to bleed down through weep line  49  and/or through the communicating, downstream parallel servo circuit components communicating with pipe  43 . Depressurized chamber  40  then allows gas pressure within regulator pressure sensing chamber  19  to move regulator sensing piston  16  and connected flow control valve  14  upward against valve seat  15 , closing off orifice  21  blocking gas flow from pipe  6  through pipe  18  to downstream gas injectors  22 . Injectors  22  may be deactivated by ECU  31 . Petrol injectors  28  then operate with fuel supplied by fuel pump  7  through fuel line  5  and rail  30 , controlled by pressure regulator  56 . 
     In a gaseous fuel only mode, petrol injectors  28  are deactivated by ECU  31  while solenoid valve  11  is energized, opening conduit  9  to allow variable hydraulic pressure to communicate with amp chamber  40  Pressure regulated gas from chamber  19  then flows through pipe  18  and rail  20  to gas injectors  22 , activated by ECU  31 . In a supercharged “dual-fuel” application, gas and liquid fuels may be injected simultaneously within throttle body  24 , as where a heat absorbing fuel such as methanol may be variably utilized with a gaseous fuel to cool the inlet fuel-air mix in order to reduce detonation and add power. This may be accomplished by selectively activating liquid injectors  28  in response to boosted air charge pressures, while gas injectors  22  and solenoid valve  11  remain continuously operative to supply the main gaseous fuel charge. Diesel dual-fuel operation employing a diesel fuel pilot charge injected into the cylinders as an ignition source, with the main gaseous fuel injected into the inlet air at throttle body  24 , can be accomplished by utilizing pump  7  as a lift pump to feed a high pressure diesel fuel injection pump and injectors (not shown) through fuel line  5 , while simultaneously utilizing all of the variable hydraulic servo components of the present embodiment to deliver a variable pressure fuel supply to gas injectors  22 . 
     Throttle by Wire 
     ECU  31  may receive fuel demand signals from sensors (not shown) that measure engine speed, and from sensors within throttle body  24  that measure manifold pressure, inlet air mass flow and temperature. Fuel tank quantity may be determined by ECU  31  from signals received from pressure/temperature sensor  23  within gas pipe  6 . Sensor  23  output can also be used for feed-forward circuitry to compensate for increased regulator output pressure that can occur with declining tank pressure. Sensor  25  located on gas rail  20 , supplies ECU  31  with pressure and temperature signals to control injector operating pulse widths, and to calculate variable supply voltages for pump  7 , valve  45  and electric regulator  56  in order to maintain variable gas pressure in rail  20  for optimal injector performance. 
     ECU  31 , receiving power demand input from an operator controlled “gas pedal” may control engine output by variably controlling gas injector rail pressure through modulation of hydraulic servo components pump  7 , valve  45  and/or electromagnetic regulator  56 . Variable servo pressure thereby produced variably actuates hydraulic amp  32  to produce in an amplified variable gas injector rail pressure from regulator  10 . Pneumatic actuator  33  or a throttle motor powered by ECU  31  then regulates inlet air flow via throttle valves  29  in response to variable fuel pressure to produce an optimum air/fuel ratio. 
     Safety 
     Referencing  FIG. 1  and  FIG. 2 , diaphragm  42  is clamped between the upper  1  and lower  3  halves of hydraulic amp  32 , and primarily serves to seal hydraulic servo pressure within chamber  40 . Annular space  44  surrounding piston-pushrod  34  is vented to a suitable place outside of the vehicle through communicating vent conduit  17 . Vent  17  with appropriate connected piping, serves to direct fuel to a safe area should diaphragm  42  or pressure sensing piston  16  within gas regulator  10  leak. In  FIG. 1 , diaphragm  42  is held against the periphery of piston crown  35  of piston-pushrod  34  by attached retaining ring  41 , represented here in cross section. Should high pressure gas from regulator  10  leak past regulator piston  16  and structure  34  into annular space  44 , excess pressure in space  44  will force structure  34  upward with diaphragm  42  and ring  41  to block orifices  52  and  48 , preventing the ingress of gas into chamber  40  and the communicating vehicle fuel system. This possibility is reduced by the venting function of conduit  17 . 
     Referencing  FIG. 2 , the function of retaining ring  41  may be supplanted by plate  64  which secures diaphragm  42  to crown  35  of piston-pushrod  34  with bolt  68 . Air pockets within chamber  40  can rise and exit through orifice  52 , eliminating the need for orifice  48  and weep line  49  shown in  FIG. 1 . A high pressure leak from regulator  10  into to space  44  will force piston-pushrod  34  upward with diaphragm  42 , causing the top surface  70  of bolt  68  to contact seating surface  50  surrounding orifice  52 , blocking the ingress of high pressure gas into chamber  40  and the communicating vehicle fuel system. In  FIG. 2 , sealing diaphragm  42  and diaphragm retaining plate  64  may be replaced with an O-ring contacting the inner wall of amp  32  positioned within an annular groove machined into the side of crown  35  (not shown), in order to seal hydraulic servo pressure within chamber  40 . Bolt  68  or a convex valve head means formed at the apex of crown  35  may block orifice  52  in the event that a high pressure gas leak from regulator  10  forces structure  34  upward against orifice  52 . 
     Alternate “Two Stage” Embodiment 
     Referring generally to  FIGS. 3-3   b , wherein like numerals represent like parts, there is depicted a two-stage embodiment, in accordance with the present invention. In order to counteract erratic low servo fluid pressure during engine idle that may result from fluctuating low pressure engine driven fluid sources or voltage driven sources subject to intermittent loads, a “two stage” embodiment of the present invention seen vertically disposed in the  FIG. 3  cross section may be employed utilizing load spring  180  positioned between regulator sensing piston  16  and piston-pushrod  134 . Referencing  FIG. 3 a   , initially low servo pressure within hydraulic amp pressure chamber  40  is applied to collar  182  and added to that of piston crown  135  through flange  184  located around the lower circumference of crown  135 . Piston-pushrod  134  and collar  182  initially move down together compressing load spring  180  against regulator piston  16  until piston-pushrod travel is limited when collar  182  contacts amp housing base  186 . This initial operational “first stage” puts spring  180  in the spring&#39;s  180  active pressure range against regulator piston  16 . No further movement of collar  182  is possible once the collar  182  bottoms on amp housing base  186 . First stage regulator output is thus determined solely by load spring  180 . Piston-pushrod crown  135 , is sealed against the inner circumference of surrounding collar  182  by O-ring  88  located within a groove around the inner circumference of collar  182 , and O-ring  90  located within a groove within the flange  184  of piston-pushrod  134 . 
     Referencing  FIG. 3 b   , “Stage two” increasing engine fuel demand causes rising servo pressure within chamber  40 , amplified solely by smaller diameter piston-pushrod crown  135  within the now stationary collar  182 , to overcome the 50 lb. counterforce of spring  180 . Piston-pushrod  134  may then continue moving downward alone within the now stationary collar  182 , exerting greater force against the load spring  180  thus causing a proportional rise in gas regulator output pressure up to and beyond the 89 lb. limit of the load spring&#39;s  180  pressure range limit in concert with engine fuel demand. 
     The description of the invention is merely exemplary in nature and, thus, variations that do not depart from the gist of the invention are intended to be within the scope of the invention. Such variations are not to be regarded as a departure from the spirit and scope of the invention.