GB2488929A

GB2488929A - Low inertia fuel pressure actuated inward opening direct injector

Info

Publication number: GB2488929A
Application number: GB1209445.4A
Authority: GB
Inventors: Vilis Ivars Lietuvietis
Original assignee: Individual
Current assignee: Individual
Priority date: 2012-03-20
Filing date: 2012-05-28
Publication date: 2012-09-12
Also published as: GB201207728D0; GB201209445D0; GB201204878D0; WO2013088428A2; WO2013088428A3

Abstract

A direct injection inwardly opening differential valve type nozzle based liquid or gaseous fuel injector with improved speed and frequency of operation by use of an ultra low inertia valve closer unit 10 positioned within the nozzle insertion tip 1of the injector, and with a valve tip 20 which is biased to the closed position by a spring 3 which also has at least part of its axial length within the nozzle insertion tip, and where the entire inertial mass of the valve closer unit 10 is preferably positioned at the tip of the injector. The injector is of benefit but not limited to, VCO and SAC nozzle type fuel injectors piezoelectrically operated from a fuel rail source. High-speed operation with negligible fuel leakage return flow of said nozzle benefits by being driven by the high-speed outward opening piezoelectric metering valve unit (29, fig.14), wherein return flow may be eliminated entirely by application of a venturi preferably inside the fuel injector.

Description

NIL INERTIA FUEL PRESSURE ACTUATED INWARD OPENING DIRECT INJECTOR

FIELD OF THE INVENTION

The present invention pertains to direct injection inward-opening commonly called hydraulic (also known as a mechanical) liquid or gaseous fuel injector based on fuel pressure actuated inward opening differential valve type nozzle improved by application of an ultra low inertia needle', offering to replace a broad variety of fuel injectors, including the currently popular VCO and SAC nozzle type fuel injectors piezoelectrically operated from a fuel rail source.

PRIOR ART PATENTS

Inward opening direct injection fuel injectors have never to the present author's knowledge enjoyed elimination of their high inertial mass valve needles. Challenging the creation of a simple and inexpensive ultra high speed precision microinjection fuel injector has been the size and mass of its valve needle, which is difficult to accelerate quickly, and once that has been accomplished at considerable expense and complexity, is impossible to stop instantly without an uncontrollable bounce, creating havoc with the injection. As a consequence, recent fuel injector technology permits the valve needle, once lifted at high speed, to go ballistic", or fly against its spring only (without a mechanical travel stop) to the extent of its energy. The price paid for an extremely fast valve opening time is an extended,,ballistic" flight of the needle against its spring which also eventually returns to its valve seat with an impact comparable to the original kinetic energy of the massive needle in its upward flight, sometimes resulting in a bounce. Such constant high impact requires large, robust valve seats to withstand the wear, which in turn demands stronger valve springs causing in turn a higher rate of wear. But more importantly, such ballistic flight consumes significant time which could be better used for precisely controlled multiple short and powerful injection pulses. Having eliminated the massive needle, the present invention has brought this technical goal into affordable reach. Valve needle bounce from its seat upon valve closure is also a known problem for the same reason as above, likewise solved by the present invention.

The only direct injection fuel injectors eliminating nearly all of the inertial mass of fuel inector needles known to the present inventor are his own previous inventions, of which one has been published as PCT/EP2OIO/061790, which unfortunately does not offer a working example of the present category of fuel injectors (namely, inward opening), but the reverse is true: The present invention adds the present category of fuel injectors to broaden the scope of application of PCT/EP2OI 0/061790.

Since the state of the art has not recognized the approach of massively reducing the length of the needle valve such that it is no longer recognizable as a,,needle', the best examples of the art are old and very simple embodiments of the very simple principle of the present invention whose simplicity matches that of the present invention, and accentuates the nature of the inventive difference between the two.

Patent US 3722801 and US 3224684, being very nearly identical patents, illustrate very clearly a simple and classic hydraulic fuel injector of the present invention category, of which the present invention is a direct improvement, dramatically illustrated. Here the long needle valve, and its operating principle, is plainly illustrated in the drwaings, whereby the inertial mass reduction advantages and nuances of the present invention, such as reduced tendency to bounce, wherein the elongated needle valve of US 3722801 is immediately recognized for its considerable elastic compressibility in its axial direction, causing it to bounce from a rigid surface in a viscous medium even when opposed by hydraulic and spring pressure, as in the present invention, wherein the bounce of an object of negligible mass would be strongly inhibited.

Patent US 4083498, FIG 3, shows a clearly different embodiment of the same operating principles of the previous two prior art examples, US 3722801 and US 3224684, wherein a small rearrangement of a few elements is apparent, but the same massive needle valve which has been taken for granted as unchangable for decades dominates.

As for other prior art, practically any recent direct injection fuel injector nozzle valve, especially of the past decade until the present time, illustrates use of a relatively massive nozzle valve needle in comparison with the present invention. A recent patent which well illustrates and analyzes the scope of the present state of the art is EP2050951A1. In the general absence of similar examples in the art to the technical approach of the present invention, this patent offers a consise broad survey of the art for general comparison. The present invention appears to offer similar advantages by entirely different means to those of EP2050951A1 in every practical respect. Although EP2050951A1 avoids entirely return flow of fuel to the fuel tank, the present invention returns a miniscule amount, consisting only of seepage past the sliding bearings in the present invention, negligible in terms of burdening the fuel pump, which is the major consideration regarding the massive return flow involved in much of the state of the art. Therefore, the present invention practically offers all of the advantages of EP2050951 Al, and suggests having some unique advantages of its own, namely high-speed performance with considerable simplicity of construction.

The relatively simply devised inertial mass reduction of the valve needle of the present invention by shortening its length so that it is no longer recognizable as a,,needle" (and for the present invention technically named the valve closer unit" unless the context calls for reference to the needle") solves a variety of technical problems involved in devising faster opening and closing speeds of the nozzle valve (which is needle-shaped in virtually all modern direct injection fuel injectors, and has therefore popularly been so named), not to mention reduced valve seat wear due to the reduction of valve closer unit impact to its seat. It is well known that faster needle operation directly contributes to lower combustion emissions, and enables more precise injection control for better combustion control for higher efficiency, better torque and power, and lower emissions. The state of the art to this end provides generally complex solutions in comparison to the solution of the present invention at substantial expense with limited results. Obviously, however, it is not the faster needle (in the present invention no longer referred to as the,,needle" but the,,valve closer unit") which creates thes benefits directly, but rather the sharper rise and fall times of the fuel flow waveform through the injector nozzle holes, which contributes to better fuel atomization and penetration into the combustion chamber of more frequent injections of shorter duration, but not neccessarily reduced fuel content. Therefore an additional, very simple and inexpensive technique invented by the present author in a separate patent application for achieving this latter effect which acts independently of the valve closer unit speed to achieve the same result is added as a desirable option to example embodiments of the present invention to make doubly sure if not doubly well that the described benefits may be achieved in at least a novel and technically useful manner, which is believed by the present author to also excel in the performance and price arenas.

The present invention applies most importantly to 7mm and 9mm diameter direct injection nozzle valves, whose nozzle body at 7mm diameter is typically from three to four such diameters in length, or about 20mm to 30mm in length. Perhaps equal in importance are the similar 9mm diameter nozzles, wherein the present invention is even easier to implement, which covers the majority of modern road vehicle propulsion. Beyond this, the simplicity of the present invention lends itself to broad unspecified application.

The invention proposes to benefit by pulse width modulated microinjections of shorter duration and higher frequency per combustion cycle than previously possible, enabling ultra refined digital micro management of the combustion cycle for superior power, torque, and emissions performance. In its lowest technology applications it promises improved performance at lower cost due to simplicity of design.

The present invention advances the state of the art by an inventive nozzle valve improvement based on the operating principle of the previous generation of fuel injector nozzle valves (often constituting the entire fuel injector) which dominated the state of the art until 2007, which were termed "hydraulic", or "mechanical" fuel injectors. The present invention approach of controlling (opening and closing) the (bounceless valve lift limited nozzle valve) ultra low inertial mass (ultra-fast and bounceless) fixed pressure opening (by valve spring bias only) purely hydraulic (also termed "mechanical") operated nozzle valve of the present invention by means of the fast hydraulic control of the outward opening directly driven piezoelectric poppet valve of the piezoelectric metering valve unit of the present invention (or any alternative equivaient) should rival the current more complicated electro-hydraulic nozzles (also referred to as "electro-hydraulic injectors, or EHI, or "sevo-hydraulic" injectors) of the state of the art (which are not limited to opening at a fixed, valve spring bias determined opening fuel pressure) not only in increasing nozzle valve speed of operation challenging the state of the art, but in reducing electrical energy requirement for the injection operation, rivaling the state of the art, as well as reducing the returned fuel flow to the fuel tank to a negligible amount, requiring no increase in the injection pump capacity over the state of the art the state of the art, all other parameters being equal.

Furthermore, an optional inventive application of the venturi principle in the present invention offers the option of totally eliminating a return flow connection to the fuel injector.

The present invention also offers benefits over the so-called "direct acting diesel common rail system" (DACR), wherein the nozzle valve needle is directly, mechanically operated by the piezoelectric stack for exceptional operating speed, which has certain aspects of superior performance over the EHI common rail injectors, wherein the present invention competes with practically every DACR advantage. This advantage resides to a large extent in inventive combination of a piezoelectric actuated outward opening mushroom valve based metering valve unit controlling its ultra low mass, highly fuel pressure change sensitive and responsive valve closer unit of the nozzle valve.

Although the bounceless, ultra-fast nozzle valve and the fuel metering valve unit of the present invention are used in combination for maximum technical effect and benefit, they can each be separately used in combination with other technologies in beneficial ways.

The present invention is adaptable to multiple fuel operation, capable of virtually simultaneous injections of multiple liquids and gases from its single nozzle valve due to its exceptional high speed, high frequency injection capability, wherein multiple metering valves or pumps may be connected to the same fuel injector nozzle valve inlet channel by use check valves. In principle, there is no limit to the number of such metered high-pressure inlet channels may be joined to the nozzle valve inlet channel, referred herein as the "combined switched flow supply channel". By this means an internal combustion engine may be fueled by gaseous fuel, liquified gaseous fuel, liquid fuels such as diesel or gasoline, or any of the previous fuels supplemented by water injection. Due to the flexibility inherent in its high frequency injection capability, which enables multiple injections of various fluids in high speed alternate injections, with the ability to nearly instantly purge the injector of a previously injected foreign liquid, insuring unadulterated alternating injections of diverse fluids, the present invention has the noteworthy capability of fueling diesel engines by compressed natural gas, wherein a pilot injection of diesel fuel provides the original ignition environnient for subsequent compressed natural gas injections, thUs maintaining the injector's lubrication both in the ignition injection prior to the gas injection, as well as in the purge injection which clears the injector of its gas content, replacing it with diesel fuel. This same capability also enables opening of the nozzle valve by a short but more powerful (higher pressure) diesel fuel injection impulse, in order to continue that same injection by an instant transition to compressed natural gas flow, without closing the valve. Thus, alternating diesel and compressed natural gas injections may be repeated many times during any given combustion cycle. It is also expected that compressed natural gas (or other such gases) should be able to open the nozzle valve of the present invention without aid of a diesel injection impulse. This is made possible by the unusually short injection pulse capability of the invention. Water, or other advantageous non-lubricating fluids may be injected by the present invention in the same manner.

DISCLOSURE OF THE INVENTION

The electrohydraulic and direct acting common rail direct injection inward-opening fuel injectors which replaced traditional hydraulic fuel injectors as the dominant fuel injector technology in 2008 have made incremental gains in fuel injector performance, but have attacked the problem at the wrong end -attempting to force an overweight and obsolete needle valve to move faster by pushing it harder. That approach avoids treating the disease by attacking the symptoms, forcing the disease to manifest itself in other places, in unexpected ways. The true prescription for a healthy fuel injector is to recognize that its central moving part is grossly overweight. In taking this approach, the present invention, in its most ambitious embodiment, applies a method of piezoelectric hydraulic stroke amplification of the nozzle valve by an indirect, and unobvious means: applying to an ultra light weight, bounce resistant valve closer unit (the successor to the obsolete,,needle valve) metering valve actuation which is amplified not by complex hydraulic pistons to amplify pressure, but whose fluid flow capacity is amplified at a fixed fuel rail pressure by simply increasing the diameter of the metering valve's outward opening mushroom valve to as large a diameter as is necessary for optimum nozzle valve flow performance at the fixed fuel rail pressure which is switched on and off (or even throttled if neccessary) by a piezoelectric actuator unit. This approach flies in the face of conventional wisdom commitment to an overweight,,valve needle".

A valve closer unit which is considerably lighter than its predecessor needle valve" does not require nearly the same amount of fuel pressure to hydraulically move it at a comparable speed, with the added advantage that it is far less challenging to stop the same valve closer unit, where a combination of fuel pressure, fuel viscosity, and the hydraulic effect of adhesion of parallel plane surfaces submerged in a viscous liquid to adhere to each other all contribute to inhibit the light-weight valve closer unit from bouncing away from its travel limiting stop -an unprecedented advantage over the state of the art lincreasing valve speed and operational frequency for multiple injections per combustion event without sacrificing injection performance. So there is a counter-intuitive piezoelectric stroke amplification phenomenon in effect in the present inveniton, whereby the quite limited stroke of which piezoelectric actuators are capable remains fixed, but the flow from the piezoelectric metering valve unit is amplified simply by increasing the diameter of its outward-opening poppet valve, supplying all the fuel flow and pressure required for optimum injection performance, while at the same time totally eliminating all back flow of fuel to the fuel tank (in the present system needed only for the miniscule seepage through its sliding valve guide bearings, and not at all for any hydraulic operation) during the 95% or more of the time when the metering valve unit is closed and not energized. The result is that the full force of the fuel rail pressure becomes available instantly by an inward opening poppet valve of small piezoelectric lift, given sufficient diameter for unrestricted flow behind it increasing the power of the fluid flow, exactly as amperage increases power at fixed voltage in an electrical circuit. It is that power which quickly lifts the nozzle valve lift piston to its optimum stop position, providing an optimum fuel pressure at the injector nozzle holes with unrestricted flow, and due to said stop allowing the nozzle valve to close much more quickly than the state of the art permits, allowing more detailed control over the combustion process for greater performance with less polluton.

The present invention pertains to fuel pressure actuated inward opening injector nozzle valves, in particular to a novel nozzle valve construction eliminating 70% or more of the inertial mass of the typical valve needle, reducing the the needle to little more than its conical tip portion, in combination with a simple but novel hydraulic valve lifting mechanism in principle based on the valve lifting principle of the old fashioned so-called,,mechanical", or fuel pump pressure injector valve activation means enabling operation of said valve (herein,,valve closer unit") by hydraulic pressure switching of its fuel supply (often preferably a fuel rail), the precise timing of which is preferably controlled by a fast piezoelectric, injection computer controlled piezoelectric metering valve unit in the fuel channel between the fuel rail and the nozzle valve.

The basic operational principle of the present invention is that of the inwardly opening differential valve type nozzle introduced in 1908 by Thornycroft of England and dominant until 2008, wherein a nozzle valve closer unit (in its popular form known as the,,pintle") of said fuel injector is operated by hydraulic pressure contained within a valve lift control chamber controlled by a pulsed fuel flow controlling (metering or dispensing) means, (for example an injection pump, or a valve connected to a fuel pressure source such as a fuel rail, or a pump, or their combination, etc.) and which may be integral and/or external to the hydraulic fuel injector; wherein the part of the valve closer unit exposed to the pressure of the valve lift control chamber (ie disposed within it) comprises a hydraulic piston (which generally functions also as a valve guide) upon whose exposed end face the pressure of the valve lift control chamber acts to urge it against the opposing force of its valve spring (specifically, upon the circular area of its piston diameter), thus tending to lift the valve open, while simultaneously the same pressure exerts a valve closing force in the opposing direction upon the smaller diameter circular valve seating and sealing area of the valve closer unit exposed to the pressure of the valve lift control chamber (ie within the circular seam line of the closed valve within said chamber). The diameter difference between the two circles results in a net,,differential" force in the direction of travel of said piston. And as soon as the valve is open sufficiently to eliminate the pressure gradient through the valve seat, said valve closing force operating at the valve seat disappears, adding a boost to the valve opeining action of the piston which is important for the reliable operation of most fuel injectors of this type.

The present invention is an ideal improvement of the Valve Covered Orifice (VCO) type nozzle (benefiting the sac nozzle type as well, in terms of speed of operation), not only in terms of speed of operation, but by its elimination of the possibility of the so-called,,valve needle" to depart radially from its central axis, causing an unsymmetrical spray pattern at low levels of valve lift, due to elimination of the,,needle valve's" high length to diameter ratio, eliminating its radial movement by providing radial valve guidance support at a minial (small) distance from the valve seat. Valve needle eccentricity has been practically the only criticism of the otherwise superior VCO nozzle type.

Unusually fast response of the fuel pressure (ie. hydraulically) actuated nozzle valve of the present invention is expected by two factors: the negligible inertial mass of the nozzle valve (plus the small inertial mass contribution of the valve spring) enabling its fast opening time, and the very small and short distance fuel volume represented by the fuel channel between the piezoelectric metering valve unit and the hydraulically operated nozzle valve, which may therefore be considered incompressible for practical purposes. The small, incompressible fuel volume of the fuel channel from the injection metering valve unit to the valve seat and valve lift piston of the present invention is unlikely of supporting and sustaining a prolonged injection flow from the injector nozzle after its piezoelectric metering valve unit closes. Also, this same fuel channel, due to its short length which is isolated above 95% of its operating time from fluid communication with the long fuel channel connected to the fuel rail, should be less susceptible to sustaining or transmitting pressure waves reaching the valve lift piston powerful enough to cause disruptive, unexpected secondary injections than state of the art designs which prefer constant fluid communication from the valve seat to the source of high pressure fuel (fuel rail, etc.). In the past, the problem of uncontrollable prolonged injection flow following the closure of the injection control, or fuel metering valve, of whatever nature whether mechanical, electomechanical, electronic, hydraulic, etc., has been termed the,,collapse phase" following the end of the electrical injection control pulse. The collapse phase is sustained and prolonged by the compressibility of the fuel in its relatively large total volume, and expansion of the large volume and large surface area ballooning elastic chambers containing that pressure, as well as by possible restricted paths of flow between sections of this total volume. The example embodiments of the present invention minimize the total volume of the fuel channel connecting the outlet of the piezoelectric metering valve unit and the bottom of the nozzle valve lifting piston, and insure that there are no restricted paths of fluid flow in said volume. In short, whatever conditions there may be for a collapse phase" in the present invention, that,,collapse phase" will be minimal in fuel volume, and minimal in terms of duration, expected to be negligible.

The electrohydraulic injectors (EHI) (also termed,Servo-Hydraulic" Injectors) which rose to dominance in the industry following 2007, increasingly employing fuel rail pressure hydraulic pressure sources and piezoelectric metering control valves for switching that hydraulic valve actuation pressure, but did so hampered by certain inefficiencies avoided by the present invention.

A drawback of EHI is that its nozzle valve control method spills considerable highly heated, volatile fuel back to the fuel tank, causing emissions and requiring a larger high pressure fuel pump. In order to create the low hydraulic pressure at the top of the nozzle valve lift piston to imbalance the high fuel rail pressure present at an opposite end of the valve lift piston (termed differential pressure) so that the the nozzle valve will hydraulically lift open, an electronic control, or metering valve (solenoid or piezoelectric) opens a valve to bleed fuel from the same fuel source which simultaneously injects the fuel through the nozzle holes. In order to bleed fuel from a supply which must simultaneously inject fuel into the combustion chamber, one must severely restrict the fuel path from the top of the valve lift piston and its piezoelectric metering bleeding valve to the fuel rail, else the fuel rail pressure will collapse. Because the nozzle valve is opened and closed in response to changing fuel pressure at the top of its nozzle valve lift piston by means of a piezoelectric (or solenoid) metering bleed valve spilling both fuel and fuel pressure, and the source of this spilled fuel pressure is a very restricted fuel channel in fluid communication with the fuel rail (restricted to avoid immediate fuel rail pressure collapse through the piezoelectric bleed valve involved in this process), this neccessary restricted fuel channel which feeds fuel to the top of the nozzle valve lift piston imposes a significant nozzle valve speed limitation on this system. All of the fule which must flow into the opening cavity above said piston must flow through said flow restriction, limiting the nozzle valve closing speed.

The present invention avoids these problems, providing instantaneous nozzle valve opening and closing maximum differential pressure differentials of the normal fuel rail pressure minus atmospheric pressure to its nozzle valve lift piston through unrestricted flow channels at a virtually negligible fuel bleed burden upon the fuel pump, wherein the entire bleed volume of fuel during an engine combustion cycle consists exclusively of the negligible fuel bleed through the high precision, close-fitting slide bearings of the system, which is subject to bleeding action only during the small fraction of the combustion cycle when the fuel injector is injecting fuel. Also, 100% of the fuel expelled from the space above the valve lift piston during valve opening is sucked back into the nozzle valve lifting cylinder cavity from whence it came by the closing action of the nozzle valve spring, resulting in a miniscule amount of fuel and no heat returned to the fuel tank, due exclusively to fuel seepage past the close tolerance sliding bearings of the fuel injector.

In contrast to the dependence of EHI injectors dependence on a process of fuel spillage and speed retardation through flow restrictors, the present invention is independent of the use of flow restrictors which add their flow delay into the nozzle valve control process, making the nozzle valve slower to respond to control signals. However, it is clear that flow restrictors may be added to the present invention for any purpose of the designer to manipulate the injection process for whatever advntage under the claims of the present invention whose nozzle valve lift piston lifting and lowering principle is independent of the use of flow restrictors.

Because the entire fuel injection system of the present invention is isolated from the fuel rail pressure by the piezoelectric metering valve unit, except during the brief injection pulses, the only mechanism or source for fuel bleed off to be returned to the fuel tank besides the tiny volume of fuel ejected from the top of the valve lift piston when it operates to open the nozzle valve. However, this same identical volume of fuel which was ejected from the top of the valve lift piston as described is immediately sucked back into the same volume from which it was expelled by action of the valve spring means of the nozzle valve, which forces the nozzle valve closed and at the same time sucks this tiny volume of fuel back into the expanding volume above the nozzle valve piston. Therefore, in theory, the present system could operate at a zero level of return fuel flow if there were no leakage losses past any of the slide bearings in the system. The net fuel return flow of the present invention is due exclusively to leakage past sliding bearings which can not normally or easily be sealed. And this is an negligible time duration of a negligible flow rate. This is in stark contrast to the popular and inefficient current state of the art ER injectors which must deliberately spill a significant quantity of fuel for each injection cycle solely for the purpose of dropping the fuel pressure in a piston chamber fed through a fuel restrictor from the fuel rail or other high pressure source, in comparison to which the normal fuel bleed through various sliding bearings is negligible. For conventional technology, but not the present invention, both loss of valuable heat from the combustion process, as well as unwanted heat pollution of the fuel which must be artificially cooled, compound the challenge of fuel returned to the tank, besides requiring a larger high pressure fuel pump.

Therefore the present invention is fully comparable to the in terms of eliminating the need for an increased capacity high pressure fuel pump to the so-called direct acting piezoelectric fuel injectors which totally eliminate the need for fuel return flow to the fuel tank, and can operate using the same fuel pump capacity as the direct acting piezoelectric fuel injectors, all other conditions, or parameters, being equal. But the present invention has its own advantages over the direct-acting piezoelectric fuel injectors beyond its fundamental advantage of potentially higher operating speeds, which is that the present invention operates most easily and conveniently as an energize to open fuel injector, whereas the direct acting fuel injector well known to te present state of the art operates as an energize to close fuel injector. Therefore, the piezoelectric element of the present invention remains not energized throughout the vast majority of the combustion cycle, whereas the direct acting fuel injector must be energized throughout the vast majority of the combination cycle, causing considerably more electrical stress and heating stress for the latter.

In contrast to this popular state of the art method, the present invention maintains a constant low hydraulic pressure at the top of its valve lift piston via an open (unswitched) or optionally switched (controlled) fluid communication with the fuel return path to the fuel tank, which ultimately, if not immediately must be atmospheric pressure at the fuel tank.

Under these conditions, the present invention now switches high fuel pressure to the bottom of the hydraulic valve lift piston from the high pressure source (typically the fuel rail) by means of its outward opening piezoelectric metering valve unit, causing the nozzle valve lift piston to rise towards the constant low pressure at its top. (Fuel injectors are conventionally drawn and illustrated with their nozzle tip pointing downwards, irregardles of the physical mounting or operational orientation of such fuel injectors).

A bonus benefit of this method is that the nozzle valve seat is only exposed to high fuel pressure while it is open, and fluid communication to the source of that fuel pressure is switched closed before the nozzle valve closes. Therefore leakage of high prssure fuel past its valve seat would be at worst quite limited if that valve seat should become imperfect through wear, debris, or other imperfections, particularly during the long intervals between combustion events. In the present invention, high fuel pressure is admitted to the nozzle valve seat only at the moment that it must lift by hydraulic action of its valve lift piston for injecton.

An advantage of using an outwardly opening piezoelectric metering valve unit having the high pressure on the outside of the valve chamber and its poppet valve, is that the high fuel pressure assists rather than opposes the valve spring pressure in maintaining said valve in a closed position, such that its valve seat sealing pressure may be much higher than in the reverse situation, wherein the valve spring pressure minus the fuel rail pressure or other source must keep the valve closed at a disadvantage in pressure. This eliminates a second potential source of fuel leakage in the system.

Direct mechanical operation of the nozzle valve by piezoelectric action has also appeared in the state of the art since 2007, but with disadvantageges of being a de-energize to inject system, requiring the nearly constant energization of the piezoelectric stack. Also, such a system is known to expose the valve seat constantly to fuel rail pressure, with the disadvantages described above for this condition. The present invention is most easily implemented as an energize to inject system, as well as eliminating the problems attending constant fluid communication with the fuel rail.

The technical effect of the peculiar "snap" action of the outward opening, mushroom valve type metering valve unit (to be described in detail) in both opening and closing, is inventively exploited in combination with the uniquely low inertial mass valve closer units of all of the embodiments of the present invention, which due to that low inertial mass are uniquely sensitive and responsive to change of applied valve actuating hydraulic pressure, to greatly exceed the previous performance limits of the state of the art for direct injection inwardly opening,,differential valve type nozzle" based fuel injectors, otherwise characterized as operating in a broad sense (allowing for inventive improvement) by the principle of,,nozzle differential ratio", often generically termed,,hydraulic" (also known as a mechanical").

But the main technical problem with the state of the art, which is solved by the present invention, remains its commitment to continue using massive pintle valves of high aspect ratio (which as such are ideal elastic kinetic energy storage devices in terms of making them ideal mechanisms for rebound, or bounce -a railroad rail dropped on its end onto a steel block will bounce much higher than that same rail dropped on its side, or sideways) which can not be stopped by a valve stop without an uncontrollable and catastrophic bounce, intolerable for fuel injector operation. This fact imposes an operational speed limit for the state of the art which is not present for the present invention.

Currently popular VCO and SAC nozzle types can both benefit by the present invention, favoring in terms of overall performance the VCO type in combination with the present author's separate (earlier) annular nozzle groove invention also illustrated herein improving nozzle pressure at low injector valve lifts without increasing fuel source (eg rail) pressure, in order to improve the performance (injection frequency, fuel volume delivery per injection, and fuel plume penetration into the combustion chamber) of the extremely short injection pulses made possible by the present invention.

Fuel injector techniques such as injection pressure amplification with or without electronic control, along with many if not most of the familiar mechanical, hydraulic, electromechanical, and electronic fuel injection control means are adaptable to the present invention. Ultrasonic modulation of the fuel pressure at the nozzle holes for improved fuel atomization at lower fuel pump or fuel rail pressures and larger nozzle hole diameters is easily adaptable to the present invention.

The present invention offers very low inertial mass enabling higher needle lift velocities, which in combination with the use of a valve travel limiting stop in combination with comparitively low susceptibility to so-called,,needle bounce, or nedle rebound from the travel stop, enables an increased frequency of precision microinjections per power stroke, offering increased injection agility, or dynamic range of computer injection control of the combustion process for efficiency and low emissions, with reduced valve seat wear.

Valve closing time is slowed by the current necessity for ballistic needle" operation as a consequence of high valve needle opening velocity due to the force of piezoelectric actuation. Advantageous use of a needle travel stop to limit valve needle lift and valve needle travel time to optimum levels must be avoided to prevent disruptive needle bounce.

Beyond a certain minimum needle lift exceeded by the ballistic needle trajectory against its valve spring, there is no increase of the rate of fuel flow from the injector due to the flow limiting cross section area of the nozzle holes at any given rail pressure. Ballistic needle technology uses about twice the valve lift needed for optimum injection. Due to its low propensity to bounce from a valve closer unit travel stop, the present invention can limit valve lift to an optmum value at a very fast opening speed (due to the low inertial valve mass), and is able to close the valve at a comparably fast valve closing speed dependent on the valve spring force, much faster and sooner than can the state of the art encumbered by needle mass which is forced to go ballistic, not amenable to use of a travel stop.

Reduced valve seat wear due to small valve inertial mass and momentum, coupled with a valve travel limiting stop, enables reduced nozzle valve seating diameter, requiring less powerful, less massive, and less costly valve springs. Alternatively, a more robust valve seat may expect an increased service life over the state of the art.

The same essential design of the present invention is adaptable for use with either a spiral valve spring in the nozzle body, or a parallel stack of very thin specially corrugated Belleville springs.

Contact of the plane top of the valve lift piston with the parallel plane top of its hydraulic cylinder as its travel stop is exploited for a bonus effect reducing still further the already greatly reduced potential for the valve of the present invention to bounce away from its travel stop. This effect is pronounced for a hugely reduced valve closer unit inertial mass and elastic axial length. This plane to plane contact, besides offering a maximum surface travel stop, minimizing wear, exploits the significant hydraulic adhesion effect of resistance of two substantially smooth and parallel surfaces immersed in a liquid to quickly separate (bounce apart). This force of separation may be adjusted to assume a range of values by reducing the area of contact of the surfaces.

If in any circumstances valve bounce might arise in the present invention, an ideally available, built-in feature will eliminate it without forseen performance penalty: the cylinder pressure vent exhausting the valve cylinder chamber at the top of the valve lift piston, which passes through the piston and out of it through a pressure resistant sliding bearing coupling to a low pressure supply pipe and low pressure fuel return channel eventually returning to the fuel tank. This cylinder pressure vent, which is bored into the valve lift piston, may ideally be provided a flow restrictor within the body of the valve lift piston or its associated valve closer unit, to dampen the motion of said piston and its associated valve closer unit.

The VCO nozzle valve was chosen as the most advantageous type to benefit by the present invention, as enhanced by an annular fuel equalization groove in the conical tip of the valve according to PCT/lB2Ol 2/051 770 also by the present inventor which groove may advantageously to all of the embodiments of the present invention be is disposed a few tens of mtcrons immediately below (in the downstream direction from) the bottom (the extreme downstream direction point) of the nozzle hole orifices in the valve seat in the valve closed position. This annular groove invention, also by the present author, increases nozzle flow at relatively low valve lift, which is quite important in combination with the inherent use of a valve travel stop positioned at precisely the optimum level for fuel flow-through. Besides reducing required minimum nozzle valve lift for a given flow rate, this groove minimizes valve opening and closing times, or injection flow rate rise and fall times,

A BRIEF DESCRIPTION OF THE DRAWINGS

Drawing FIG I illustrates the present nozzle valve invention in its closed position, embodying a conical valve tip section of conventional size and shape for 7mm VCO nozzle insertion tip diameter automotive fuel injectors, maintaining its 7mm diameter to a typical nozzle insertion tip length of 3 to 4 times that diameter (20 to 30mm), housing a typical 4mm diameter insertion tip bore, conventionally for the valve needle and its fuel passage, but exploited in the present invention for housing the coil valve spring means, from which point downward features depart significantly from the typical VCO design which is displayed for convenience of reference as FIG 18.

Drawing FIG 2 illustrates the nozzle valve embodiment of FIG un its valve open position.

Drawing FIG 3 illustrates the nozzle valve embodiment of FIG I having a reduced valve seating diameter, which is made possible by the reduced valve seat wear of the reduced inertial mass of the valve, and offers some performance advantages over the embodiment of FIG 1, along with an alternative robust attachment geometry for the coil valve spring base plate and its axial fluid channel coupling slide bearing, along with illustration of a thickening of the end of the insertion tip bore of FIG 3 of the nozzle tip, increasing its fuel pressure capacity.

Drawing FIG 4 illustrates the nozzle valve embodiment of FIG 3 in its valve open position.

Drawing FIG 5 illustrates a complete, self-contained fuel injector based upon the nozzle valve embodiment of FIG 3 wherein, besides new details of the low pressure supply pipe and the nozzle unit, an inventive piezoelectric activated metering valve unit of particular benefit to the operation of the present nozzle valve invention is introduced.

Drawing FIG 6 shows an enlargement of the lower half of the fuel injector of FIG 6, emphasizing the interrelationship of the nozzle unit and the metering valve unit.

Drawing FIG 6A is like FIG 6, showing how a shorter coil valve spring means is simply accommodated within the insertion tip bore of the nozzle insertion tip.

Drawing FIG 7 shows an enlargement of the piezoelectric metering valve unit of FIG 5 and 6 in the closed position, illustrating a valve diameter of normal size for a normal fuel injector.

Drawing FIG 7A shows the mushroom valve detail of FIG 7 having an enlarged diameter which would considerably exceed the the needs of a normal fuel injector.

Drawings FIG 8 and 8A show the piezoelectric metering valves of FIG 7 and 7A in the open position.

Drawing FIG 9 shows in cross section, the hydraulic cylinder part of the nozzle tip hydraulic cylinder group, which would house the valve lift piston of the valve closer unit of the nozzle valve. The 36 black dots distributed in a circle within the cross section illustrate a maximum number of a maximum size of nozzle tip hydraulic cylinder group fuel channels routed around the valve lift piston 9, providing 12 times the flow capacity of the nozzle holes (demonstrating abundant fuel flow sufficiency of this geometry), being an important aspect of the present invention.

Drawing FIG 10 illustrates an analogous embodiment to that of FIG 3 wherein a parallel stack of Belleville springs replaces the coil valve spring means.

Drawing FIG 11 illustrates the nozzle valve of FIG 13 in the open position.

Drawing FIG 12 illustrates the Belleville valve spring means based nozzle valve of FIG 10 and 11 incorporated in a complete fuel injector, analogous to that of FIG 5, as depicted in its enlargement of FIG 6.

Drawing FIG 13 illustrates an embodiment having essentially the same parts and structure as that of FIG 3, but differing in that the high and low pressure channels supplying the nozzle valve have been connected in reverse from that of FIG 3, requiring a remarkably simple modification of the internal fuel channeling to the two valve lifting chambers involved, the hydraulic valve lift control chamber below the valve lift piston, and the valve cylinder chamber above the valve lift piston.

Drawing FIG 13A illustrates a simple variant of the embodiment of FIG 13, wherein the high pressure fuel is not introduced into the hydraulic valve lift control chamber above the valve seat first, as is the conventional practice for a direct injection inwardly opening differential valve type nozzle" based fuel injector, otherwise characterized as operating in a broad sense (allowing for inventive improvement) by the principle of nozzle differential ratio", often generically termed,,hydraulic" (also known as a,,mechanical"). Instead, the a fuel is introduced to proximity with the nozzle holes via a single orifice in the axial center of the tip of the moving (dynamic) conical valve tip section, which has the advantage that the pressure of this fuel acts as if the valve were an outward opening instead of an inward opening valve, and that upon the full cylinder diameter of the valve lift piston, making this valve much easier to open than any alternative lacking this axial fuel entry point.

Drawing FIG 13B illustrates an embodiment which is a combination of those of FIG 13 and FIG 13A. The performance of this embodiment is expected to be similar to that of FIG 13A.

Drawing FIG 14 illustrates the embodiment of FIG 13 as incorporated within a complete fuel injector, showing how the fuel connections are arranged in reverse from those of the very similar embodiment of FIG 3.

Drawing FIG 15 illustrates an embodiment similar to that of 13A, wherein need for a return fuel flow connection to the fuel tank (or a similar alternative) is eliminated. Instead of providing a source of low pressure, as conventionally by connection to the fuel tank, above the valve lift piston for lifting of the valve lift piston from below it by a source of high pressure, a venturi disposed within the body of the valve closer unit, and more specifically within the valve lift piston of the valve closer unit, is provided. In the present embodiment, even the slender part of the axial fluid pipe normally carrying the high pressure switched flow supply channel to the nozzle valve is eliminated, allowing the entire insertion tip bore housing the valve spring means to channel this switched flow supply channel for the nozzle valve. However, retention of this slender part of the axial fluid pipe and its sliding bearing connection to the valve closer unit may be retained, and in some circumstances, may be advantageous.

Drawing FIG ISA presents an enlarged view of the valve closer unit venturi and associated elements intended to optimize the performance of this embodiment.

Drawing FIG 15B illustrates an increase of the nozzle valve sealing contact ring diameter for the purpose of providing an initial hydraulic valve lifting advantage (if needed). The original lifting force is provided by the pressure from the nozzle tip acting upon the area of this circle, in opposition to the valve spring pressure, the injection pressure acting upon the bore of the piston coupling and guide pipe bearing, and the pressure within the valve cylinder chamber acted upon by the venturi.

Drawing FIG 16 illustrates the embodiment of FIG 15 as incorporated within a complete fuel injector, showing how the fuel connection is like that arranged in the embodiment of FIG 3, with the exception that the low pressure supply channel has been eliminated.

Drawing FIG 17 illustrates a Belleville spring valve spring means embodiment similar to that of FIG 10, with the exception that a venturi is employed to provide the source of low pressure required for the nozzle valve hydraulic piston to operate, in principle like that used in the embodiment of FIG 15 and FIG 16, but eliminating the spiral valve spring means. In addition, as shown in dotted lines immediately above the nozzle valve mechanism, is shown an alternate positioning and arrangement for the venturi, which may be used in parallel with, simultaneousiy with the first venturi.

Drawing FIG 18 illustrates the embodiment of FIG 15 without its venturi, showing the possibility of nozzle valve operation even without benefit of the venturi of fFIG 15.

Drawing FIG 19 illustrates an embodiment hydraulically (operationally) identical to that of FIG 18, stripped of its superfluous elements.

Drawing FIG 19A Ilustrates an increase of the nozzle valve sealing contact ring diameter for the purpose of providing an initial hydraulic valve lifting advantage (if needed).

Drawing FIG 20 illustrates details of the nozzle hole pressure equalizing groove which is of significant benefit to all of the embodiments of the present invention.

Drawing FIG 20A illustrates the precise coincidence of the upper edge of the nozzle hole pressure equalizing groove with the lower edge of the inlet aperture of the nozzle hole prior to the rouding of its aperture edge by hydro-erosion grinding (shown in dashed lines).

Drawing FIG 20B illustrates FIG 20A following the hydro-erosion grinding operation, showing a deliberate gap connecting the nozzle hole with the nozzle hole pressure equalizing groove.

Drawing FIG 21 illustrates in principle by the example of a single embodiment the method by which muli-fuel (gaseous or liquid) operation, or introduction of water injection, or any other fluid useful for management of the combustion process, may be applied to all of the embodiments of the present invention. FIG 21 illustrates specifically an embodiment which enables running a diesel engine on compressed natural gas, wherein the gas ignition is provided by a small pilot injection diesel ignition, followed by a main injection of compressed natural gas.

Drawing FIG 21A illustrates the check valve used to enable multi-fluid injections for any combustion cycle by the present invention.

Drawing FIG 18 22 illustrates a conventional VCO nozzle for comparison (valve open position). Although the drawing is truncated, the length of such a nozzle at the nominal diameter (two popular automotive standard diameters are 7mm and 9mm) is typically from 3 to 4 nozzle diameters in length.

DETAILED DESCRIPTION OF EXAMPLES OF EMBODIMENTS

The present invention is further described by way of generally sequential reference to examples of embodiments illustrated in the accompanying drawings, wherein the relative proportions and scale of the drawings is referenced to ISO standard 7mm and 11 mm diameter fuel injctor nozzles.

In the following descriptions of drawings, generically identical in form and function details or elements of diverse assemblies within a given embodiment, or generically identical details or elements of diverse embodiments, are designated with identical reference numbers. The purpose of this consolidation is to emphasize the identical function in different locations of the invention of various forms and applications of a given structure.

Drawing FIG I illustrates the present nozzle valve invention in its closed position, housed in a typical 7mm diameter VCO nozzle insertion tip I (compare with the standard VCO nozzle of FIG 19) of length over 2 times, and typically from 3 to 4 times its diameter (maintaing that diameter), of normally 7mm outer diameter and 20mm to 30mm length at that diameter, housing a typical 4mm diameter insertion tip bore 2, from which point in the upstream direction (upward in the drawings) features begin to depart significantly from the typical VCO design of FIG 19. To the best knowledge of the present author, no direct injection VCO nozzle (or direct injection fuel in jector) has previously housed its coil nozzle valve spring in its insertion tip bore 2, herein designated as valve spring means 3. The insertion tip bore 2 also serves the function of being a high pressure conduit for fuel to be injected as well as serving as a relatively incompressible hydraulic valve lifting medium working against the valve spring means 3.

For discussion of features of FIG I prominent in the axial cross section view of nozzle tip hydraulic cylinder group 4, and especially its lower hydraulic chamber external channels 5A, please refer to the description of the radial cross section of the nozzle tip hydraulic cylinder group 4 presented in a separate drawing, FIG 9, which elaborates upon this part of FIG 1. Both the axial (FIG 1) and the radial (FIG 9) cross sections must be seen together to understand their importance.

The present invention operates as an inwardly opening differential valve type nozzle, operating on the principle of nozzle differential ratio", injecting through nozzle holes 6 which are opened for injection by action of fuel pressure in hydraulic valve lift control chamber 7 whose high pressure collapses the volume of the valve cylinder chamber 8 (cf FIG 2) by lifting into that volume the valve lift piston 9 of the valve closer unit 10 by virtue of low pressure (typically sourced from the fuel tank) in the valve cylinder chamber 8.

More specifically, a nozzle valve closer unit 10 (popularly known as the,,pintle") of said fuel injector is operated by hydraulic pressure contained within a hydraulic valve lift control chamber 7 controlled by a pulsed fuel flow controlling (metering or dispensing) means, (for example an injection pump, or a valve connected to a fuel pressure source such as a fuel rail, or a pump, or their combination, etc.) and which may be integral and/or external to the hydraulic fuel injector; wherein the surface of said valve closure unit comprises under common fluid communication (pressure) within said valve lift control chamber 7 the exposed (to said pressure) circular interface boundary of an annular valve sealing surface or line (termed a first circle) sealing against said fluid communication with a nozzle hole 6 disposed downstream from said annular sealing surface; and also the exposed (to said pressure) circular interlace boundary of a cylindrical valve lift piston 9 in its sliding pressure resistant cylinder, termed a second circle, disposed upstream (farther from said nozzle hole 6) from said annular valve sealing surface, wherein the area of said first circle is smaller than the area of said second circle, whereby said pressure upon the common surface of valve closure unit 10 acting within both circles establishes an unbalanced (also known as,,differential') force upon valve closer unit 10 in the direction from said smaller to said larger circle, which lifts the valve closer unit 10 axially against the closing force of its valve spring means 3 so long as a hydraulic valve lift control chamber 7 pressure exceeding the nozzle valve closing pressure, resisting both the valve spring force and the fluid outflow through the nozzle hole(s) 6, is maintained.

In order for the valve lifting function to work, the volume of incompressible fuel contained within the valve cylinder chamber 8 (constantly vented to nearly atmospheric pressure) located at the top of the valve lift piston 9 of valve closer unit 10, namely immediately above the valve lift piston top 11, which must be radially symmetric about the valve closer unit 10 axis, preferably planar as depicted, but optionally spherical, etc., must be discharged as its plane surface rises to meet and contact its parallel plane surface (in general radially symmetric and congruent) of cylinder top 12 (which simultaneously acts as the valve lift piston 9 travel limiting stop), collapsing the volume of valve cylinder chamber 8 as its incompressible fuel content is temporarily discharged (and later sucked back into the valve cylinder chamber 8 on the return stroke by the force of valve spring means 3) into the valve closer unit fluid channel 13, a channel usually at atmospheric pressure (or leading to atmospheric pressure), typically leading back to the fuel tank. The hydraulic pressure lifting the valve closer unit 10 works against the valve spring means 3, which exerts force against the coil valve spring base plate 14 of the spring base & piston channel unit 14A, which is provided a valve closer inner flow hole l5Ato aid downstream conduit of fuel (depicted here as a specific hole, but which may be a surface groove or arbitrary hole, or combination thereof), which spring base & piston channel unit 14A has a tubular extension axial fluid channel coupling slide bearing 16 engaging by a sliding bearing interface which is pressure (leakage) resistant the axial fluid pipe 17 conducting out of the nozzle, and which spring base & piston channel unit 14A is rotationally attached to the piston coupling and guide pipe 18 by an inner, recessed sleeve bearing preferably also comprising along with valve closer unit fluid channel 13 an end bearing surface to bear and transmit to valve closer unit 10 at least in part the force of the valve spring means 3, which piston coupling and guide pipe 18 passes through the cylinder top 12 and further through a pressure resistant sliding close fitting piston coupling and guide pipe bearing 19 passing through the top of nozzle tip hydraulic cylinder group 4. This bearing, besides its fluid sealing and dynamic valve spring means 3 coupling functions, augments the valve seating axial alignment and centering function of the close sliding fit of the valve lift piston 9 within its cylinder bore in the nozzle tip hydraulic cylinder group 4, which is also the periphery of the valve cylinder chamber 8. The piston coupling and guide pipe 18 therefore augments the alignment and centering of the valve tip section 20 (reducing the common problem of singly-guided VCO nozzle valve needles drifting off of their axis, causing an asymmetric spray patterns from their ring of nozzle holes 6), and more specifically axially centering of the lower portion of the valve tip section 20, which is the valve tip congruent seating region 21, with respect to the valve seat congruent seating region 22, which comprises the valve seat part of the nozzle insertion tip 1.

The axial fluid pipe 17 serves the valve lift piston 9 action first by its principal purpose at the top of valve lift piston 9 of unbalancing the high pressure injection pulses introduced into the hydraulic valve lift control chamber 7 at the bottom of valve lift piston 9 to urge its upward motion against the valve spring means 3; second by providing a low pressure (possibly atmospheric) space for intermittent storage of the volume of fuel expelled from the valve cylinder chamber 8 by the rise of valve closer piston 9, and sucked back into the the valve cylinder chamber 8 by the descent of valve closer piston 9; and third by providing a path of return flow to the fuel tank of the very minute seepage flow past the several closely fitted, leakage resistant sliding bearings of the valve closer unit 10 subject to a high fuel pressure differential. Due to this net minimal flow rate through the axial fluid pipe 17, cooling of the returned fuel is not expected to be the technical problem it is to the state of the art electro-hydraulic fuel injectors, and neither is the loss of valuable heat energy from the injector nozzle, where this heat may assist vaporization of the fuel and the general combustion process.

It should be recognized that the optimal connection of the axial fluid pipe 17 of the present invention by means of the axial fluid channel coupling slide bearing 16 for the purpose in the present embodiment of channeling low pressure originating in the fuel tank for the purpose of providing an unbalanced pressure force to move the valve lift piston 9 by means of the high pressure fuel pulses present in the insertion tip bore 2 channeled via the lower hydraulic chamber external channels 5A to the hydraulic valve lift control chamber 7, is not its oniy advantageous arrangement. Specifically the long, slender section of the axial fluid pipe 17 which passes inside the coils of the coil valve spring means 3, could have been rigidly attached to the valve closer unit 10 at the point where the axial fluid channel coupling slide bearing 16 is optimally attached as drawn, thus replacing it in an inferior arrangement which is nevertheless advantageous by comparison with the state of the art in providing a less massive so-called "valve needle". In such a less effective arrangement (which is not illustrated by a drawing, but may be visualized by reference to FIG 5), due to its unwelcome increasing of the inertial mass of the valve closer unit 10, the increased diameter upstream (upper as depicted in the orientation of the drawings, eg FIG 5) pad of the axial fluid pipe 17 axially beyond the region of the nozzle insertion tip 1, remains unchanged from that depicted in drawing in FIG 5, etc, upwards from the shoulder of its contact with the spring pressure adjustment tube 46 at the axial location of the axial fluid pipe spiral flow transfer annular cavity 26B, specifically upwards to its boundary with the metering valve unit 29.

However in this modified and inferior arrangement, the slender part of the axial fluid pipe 17 now rigidly attached to the upper end of valve closer unit 10 will now penetrate the former larger diameter portion of the axial fluid pipe 17 at its lower end, at its former annular shoulder supporting the the spring pressure adjustment tube 46, which in its new configuration continues to support said spring pressure adjustment tube 46, functioning exactly as previously, with the only difference being that the slender part of the axial fluid pipe 17 now penetrates into it by an outer close fitting sliding bearing contact, to accommodate the axial movement of the valve closer unit 10 now enlarged to include the slender part of what was formerly the axial fluid pipe 17. Even this inferior modified arrangement could retain some of the benefit of the present invention, and is therefore claimed by the present patent, but is not depicted in a drawings of the various applicable embodiments.

If in any circumstances the problem of valve bounce might arise in the present invention, placing a valve motion damping viscous flow restrictor in the valve closer unit fluid channel 13, exhausting the valve cylinder chamber at the top of the valve lift piston 9, which passes through said piston 9 and out of it through a pressure resistant axial fluid channel coupling slide bearing 16 leading eventually back to the fuel tank, will eliminate it. Such a flow restrictor, or simply a diameter reduction of valve closer unit fluid channel 13, ideally within the body of the valve lift piston 9 or its associated valve closer unit 10, or placed possibly further in the fluid communication path back to the fuel tank, to dampen the motion of said piston 9 and its associated valve closer unit 10 to the degree that problematic valve bounce is reduced to any desired degree.

Note that because the piston coupling and guide pipe 18 is mechanically captured by its piston coupling and guide pipe bearing 19 due to elemnts of larger diameter attached to said coupling shaft 18 both above and below said guide bearing 19, assembly of the nozzle tip hydraulic cylinder group 4 with at least the spring base & piston channel unit 14A requires in the present embodiment separation of said spring base & piston channel unit 14A from valve lift piston 9. Such separation can be accomplished in various ways. First is the method shown in drawings FIG I and 2, wherein spring base & piston channel unit 14A is rotationally joined by a slide bearing connection to the piston coupling and guide pipe bearing 19, to allow the nozzle valve tip section 20 to rotate in its seat to insure radially symmetrical wear. An alternative assembly method is depicted in FIG 3 and FIG 4 as spring base & piston channel unit 14B, wherein coil valve spring base plate 14, axial fluid channel coupling slide bearing 16, and piston coupling and guide pipe 18 are a rigid unit made of one peice, which is rotationally fitted in a relatively deep slide bearing hole bored in the center of valve lift piston 9 (or deeper). Yet a third variant, spring base & piston channel unit 14C is depicted in the embodiment of FIG 15, wherein high pressure fuel is supplied via the valve closer unit fluid channel 13, rather than venting low pressure as in the present embodiment, and no return flow venting is required by application of a venturi.

Yet a further variants, employing Belleville springs, omit this item altogether. Alternative combinations of features of these and other variants of the present document are not shown, as would be evident to the practitioner of the art, which are claimed by the present invention.

Note that the valve tip congruent seating region 21, in its relationship with respect to the valve seat congruent seating region 22, which comprises the valve seat part of the nozzle insertion tip 1, does not specify within either of the location of one or more neccessary annular hInes or bands of valve sealilng contact, which has become a field of considerable creative engineering variation. This is because in the VCO nozzles of the example embodiments are employed to advantage, and always in combination where both are present as is generally the case, two quallities or types of fluid sealing: absolute and partial. The dominant fluid seating type is the absolute pressure tight annular line or band of valve sealing contact between the frustoconical valve seat and valve closer parts coming into contact which frustoconical geometries have slightly differing cone angles, forcing a line contact between them, rather than the entire broader surface to surface contact which may otherwise be possible. This line contact is in the most elementary and oldest form of the VCO nozzle the upstream boundary of the two named seating surfaces, disposed upstream of the nozzle holes 6, providing a pressure tight seal. In contrast to and in combination with this perfect high pressure tight seal is the highly effective leakage resistant imperfect seal of the VCO nozzle comprising a valve gap in regard to the trapped fuel volume downstream of the said pressure tight seal, and entirely dependent upon said pressure tight seal, but nevertheless exposed to the nozzle holes 6 which provide fluid io communication with the combustion chamber. Leakage from this trapped volume due to the small angular difference between the congruent frustoconical sections of the valve seat is greatly inhibited for the following reasons: first, the fuel within the trapped volume is nearly incompressible, and it can not go anywhere; second, where it is exposed to the orifices of the nozzle holes 6 it is confined within an extremely narrow annular channel formed within the described slightly divergent conical surfaces which resists the flow of fluid; third, it is surrounded by relatively massive adjacent steel parts which prevent or at least limit its high temperature flash boililng and vaporization; and last but not least, however it works, practice has proven that it works as demonstrated in the well-known superiority of the VCO nozzle over the sac-type nozzle in terms of hydrocarbon emissions, wherein the trapped fuel volume of the sac-type nozzle is not confined narrowly between adjacent conical steel surfaces, but is closely and densely surrounded by nozzle hole 6 apertures, creating vastly greater exposure to the influence of the combustion chamber than in the case of the VCO nozzle even though the total exposure area of the nozzle hole 6 apertures is equal in the two cases.

All of the example embodiments of the present drawing are drawn to a scale which ideally depicts a 7mm outside diameter nozzle insertion tip 1, and may also accurately represent a 9mm nozzle insertion tip 1. But the context of the discussion for all of the example embodiments is in terms of the 7mm nozzle insertion tip 1. Of course, the invention itself is not limited to any specific size or scale.

The VCO nozzle hole 6 pressure equalizing groove 23, being 160 microns wide and 100 microns deep as represented in all of the 7mm diameter nozzle insertion tip lexample embodiments of the present drawing, intended to raise the fuel pressure at the nozzle holes 6, drawn as 120 microns diameter in all of the example embodiments of the present invention, is beneficial though not essential for the operation of the present invention, but its potential benefit is considerable in just the area of special advntage of the present invention, which is its superior valve speed potential, and especially its potential to exploit this superior valve speed for the creation of uncommonly short injections which are fully penetrative and fully vaporized, since short injections of the state of the art are plagued by poor penetration and less than optimal vaporization due to low nozzle hole 6 pressures during the initial low levels of valve lift. Thus the nozzle hole pressure equalizing groove 23 increases the fuel pressure at the nozzle holes 6 during low levels of valve lift, enabling high fuel flow at low valve lift. This feature is important in combination with the valve lift limiting stop of the present invention, insuring that valve lift travel ends at the optimum value of valve lift, wherein that optimum leve is reduced by the flow-through enhancing ability of the nozzle hole pressure equalizing groove 23.

In this connection, it will be observed by comparison of the following embodiment illustrations that the top of the valve cylinder chamber 8 acts as a travel stop, or maximum travel limit for the valve lift piston 9, wherein the overall effectiveness of the fuel injector is increased by the combination of this limitation of the valve lift in combination with the stated benefits of the nozzle hole pressure equalizing groove 23, accomplishing a greater amount of fuel injection in a shorter amount of time, with less valve travel and wear and tear than is possible by the state of the art.

Here a valuable attribute of the present invention can be best appreciated, in its promising ability to avoid or practically limit so-called needle bounce" in modern conventional direct injection fuel injectors which employ relatively massive and long needle valves, which in order to achieve their fast opening times must by ingenious and expensive means be accelerated from their valve seats abruptly and forcefully, imparting to the needle valve large kinetic energy which causes the needle to rebound forcefully from any mechanical stop blocking its ballistic travel away from the valve seat and against the valve spring.

Such rebounding causes at least pressure waves which disturb the injection quality, if not premature needle closing. Therefore current technology permits the fuel injection nedles to assume a ballistic" mode of travel against the valve spring, which completely avoids the use (and advantages) of a mechanical travel stop.

But in the present invention, due to the extreme low inertial mass of the valve closer unit 10, in combination with valve closing valve spring means 3 which may be comparable to the normal state of the art, due to its need to resist combustion chamber pressures (being also somewhat dependent on the valve seating diameter), the valve closer unit 10 of the present invention would have a very small fraction of the rebound energy of a state of the art needle valve, which very small rebound energy would be opposed by the very high hydraulic force which lifts the needle. The inertial motion of the spring means 3 is partly involved in this valve lift, but this motion is linearly distributed along its length, and the spring is unlikely to lift from its contact with the spring base & piston channel unit 14A when the valve closer unit 10 is suddenly stopped by colision with cylinder top 12 (or such lift would be of no practical effect against the fuel rail pressure on the valve lift piston 9), resulting in a dissipating pressure wave being propagated in time and distance upward along the spring, which will not disturb the valve closer unit 10 being forced against its stop by the full and immense fuel rail pressure. Also, the present valve closer unit 10 would have virtually zero change in lengt as a consequence of striking its rigid travel stop, whereas a Jong injector needle wouid exhibit considerable elastic deformation, which would spring back, affecting the fluid volume during injection, or even worse, propelling the needle prematurely against its seat.

Important in this process is when the plane (or otherwise nominally radially symmetric congruent surface) valve lift piston top 11 meets the plane (or otherwise radially symmetric congruent surface) top of the valve cylinder chamber 8, all liquid is expelled from between the two plane (or otherwise radially symmetric and congruent) surfaces in full contact of their nominal surfaces (excluding contact at surface locations containing apertures or grooves opening into and below the nominal surface serving special purposes, such as venting holes), which in the experience of most people who have pressed the plane faces of two sheets of glass or other such near perfect surfaces together under water (or merely stacked wet sheets of glass together), establishes a very significant resistance to pull them apart, which in the case of the low mass valve closer unit 10 (which may be further lightened by boring holes, etc.) would significantly hinder any tendency of the exceptionally low-mass valve closer unit 10 to rebound from its stop, the cylinder top 12.

And annular corrugation of the two congruent mating surfaces, valve lift piston top 11 and cylinder top 12, involved in the valve travel stop would not only increase their contacting surface area, but would increase the path of valve lift piston 9 motion damping and retarding viscous fluid flow through their narrow dynamically expanding or contracting gap, as well as removing straight paths of fluid flow, further increasing the resistance to viscous flow. And furthermore, the amplitude of the corrugation increases the duration of the cyclical period of valve closer unit 10 which is under the influence of restricted flow into and out from the narrow gap between said two congruent mating surfaces. The greater said corrugation amplitude, the longer is the period of valve lift piston 9 during a condition of reduced clearance between said two congruent mating surfaces. Excessive such corrugation amplitude will ultimately immobilize the valve closer unit 10.

Moreover the contact of these two surfaces during the period of fuel injection insures a valve sealing contact which seals the orifice(s) and apertures in the surface of these two surfaces which are neccessry to collapse and to expand the fluid volume of the valve cylinder chamber 8. Such a valve sealing contact is beneficial to the fuel injector operation because it stops possible leakage of fuel through the sliding gap between the valve lift piston 9 and its cylinder back to the fuel tank. When applied in combination with the inventive use of the pulsed fuel flow controlling (metering or dispensing) means of FIG 7 and FIG 8, which insure that no return flow leakage is possible during the period during the period when fuel injection is not taking place, these two limitations on return flow to the fuel tank virtually eliminate all net return flow back to the fuel tank, soliving an important technicai problem and important goal of much of the state of the art.

This inherent absence of nozzle valve rebound in the present invention can be exploited to considerable advantage for the purpose of achieving computer controlled extremely precise and short injection pulses of maximum flow rate (having very sharp rise and fall times to and from the maximum flow rate of the injector), whch is a goal of the art which has thus far been elusive. Such ability would permit very fine and detailed control of the combustion cycle by applying a large number of injection pulses of any duration, but especially of short duration, to digitally control the combustion process for any given cycle, which depends on many engine parameters such as engine speed and load, acceleration, ambient conditions such as temperature and pressure, et cetera. The possibilities are extensive and complex.

And a major impediment to such short injection pulses by VCO nozzles is that most of the injections from a VCO nozzle occur under conditions of throttling by the valve needle.

Throttling means that the valve needle has not reached its optimal elevation for fuel injection. This throttling of state of the art VCO nozzles (unimproved by the present author's present invention substantially increasing the valve opening and closing speeds which greatly limit the duration of throttling, and augmented by the present author's separately patented invention, previously introduced as nozzle hole pressure equalizing groove 23, which virtually eliminates throttling, regardless of valve speed) more than for sac type nozzles discourages their application to meet high performance demands as have just been described, through the creation of very sharp and square very short injection flow pulses, wherein the nozzle holes, and no other flow restriction limits the injection flow. The thesis crua.net/research/FDFs/Crua_FhD_Thesis_Ch4.pdf reports on p. 81 that for short VCO injection pulses of about I ms, most of the injection occurs under conditions of throttling, in contrast to normal pulses of 3 or 4 ms where throttling is not so predominant.

P. 83 of the same document reports "Influence of injection pressure. Figure 4-14 shows the effects of increasing fuel rail pressure on a VCO nozzle spray... The injection pressure was found to have a significant effect on the rate of penetration of the liquid core, for a constant injected fuel quantity (injection time was varied...). As expected, higher injection pressures produce fully developed sprays in a shorter time, thus improving the vaporisation process as the surface area of the liquid core increases." The ability of the nozzle hole pressure equalizing groove 23 to increase fuel pressure at the nozzle holes 6 at low levels of valve lift therefore makes the usefulness of such a groove vital to the ability of the present invention to produce unusually abrupt (suddenly vertical) injection flow rate rise and fall times, due to the negligible inertia of its valve closer unit 10, and anticipated absence of tendency to bounce from a rigid valve travel stop or limit. Since fuel flow does not increase with increasing needle lift beyond a quite limited lift distance (due to flow throttling by the nozzle holes 6), the current practice of ballistic needle travel for heavy needles severely limits the ability to create ultra short injection pulses, which indicates that use of a needle travel stop can be used to great advantage with the present invention, to narrow the time resolution of valve control substantially for both needle opening times and needle closing times and for the interval between the two.

In the present invention the valve closer unit 10 is available for action (control) throughout a larger proportion of the combustion event than in the conventional use of heavy ballistic needle which spends a great deal of time flying in a ballistic mode which Is not subject to interruption by a need to change the valve closer unit (needle) position.

Therefore both the axial position of the valve closer unit travel stop (ie cylinder top 12) and the nozzle hole pressure equalizing groove 23 used in combination such that the nozzle hole pressure equalizing groove 23 is positioned to provide near maximum fuel pressure at the nozzle holes when the valve lift piston top 11 is in contact with (stopped by) the cylinder top 12 offers to provide otherwise difficult to achieve (very fast and high repetition frequency) needle control response with improved combustion performance at reduced cost. The upper edge of the nozzle hole pressure equalizing groove 23 determines how early the influence of said groove will begin to sharply increase the fuel pressure at the inner orifices of the nozzle holes 6, as said groove 23 approaches sufficient fluid communication proximity to said nozzle holes 6. The width of the nozzle hole pressure equalizing groove 23 determines the duration of its pressure increasing effect. And the axial position of the valve closer piston 9 stop, the cylinder top 12, can advantageously stop the rise of the nozzle hole pressure equalizing groove 23 at a point where its pressure increasing effect at the input orifices of the nozzle holes 6 reaches a desired maximum level of fuel flow, which may be conditioned by many factors, such as fuel atomization, etc. By such coordination of said groove 23 axial position in relation to the axial position of the travel stop, the cylinder top 12, the width of the nozzle hole pressure equalizing groove 23 may be made optimum -not too narrow and not to wide, and with an optimum depth determining the flow rate.

Drawing FIG 2 illustrates the embodiment of drawing FIG I in the nozzle valve open position.

The fuel path down the insertion tip bore 2 (ie in the downstream direction) is comprised significantly, but in most circumstances not exclusively, of a nozzle outer spiral fuel channel 24A and an valve closer inner spiral fuel channel 24A. In the abnormal and idealized circumstance wherein the outer radial extremity of the coil valve spring means 3 contacts its adjacent cylindrical wall of insertion tip bore 2 by a fluid tight contact, and/or the inner radial extremity of the coil valve spring means 3 contacts its adjacent cylindrical axial fluid pipe 17 by a fluid tight contact, and/or the coil valve spring means 3 is fully compressed such that the turns of the coil are in fluid tight contact, preventing fluid communication between the outside of the coil valve spring means 3 and its inside, in such an unusual circumstance the coil valve spring means acts as a hollow cylinder allowing fuel to flow downstream both inside and outside its cylindrical wall, but not through the wall. The purpose of this exercise is to describe the geometry for the fuel entry and exit orifices at the support surfaces of the coil valve spring means 3 under an imaginary (not realistic) worst-case scenario where the coil spring is fully compressed, preventing fuel flow from the inside to the outside of the coils. We will consider the fuel flow-through requirement pertaining to the valve closer inner spiral fuel channel 24A, the inner flow channel of the coil valve spring means 3 at its lower supporting end, the results of which will apply likewise to its upper end.

Under this worst-case scenario, (whereof the present drawings are not representative, offering significant space both between successive turns of the valve spring means 3 as well as significant space between the inner surface of the valve spring means 3 and which is apparent from the proportions of FIG I and FIG 2, two distinct spiral fuel channels are evident: a nozzle outer spiral fuel channel 24A and an nozzle inner spiral fuel channel 25A.

Spiral valve springs most commonly have their supporting end surfaces ground flat to rest in plane to plane contact with its supporting surface, the coil valve spring base plate 14 of the spring base & piston channel unit 14A. The two-dimensional profile of this flat-ground coil surface (the surface of mutual contact between the planes) resembles a nearly circular crescent moon shape (rather like the crescent of an eccentric lunar eclipse) with its two pointed ends in contact: the widest part of this crescent shape is directly opposite its two pointed ends in contact through the axis of the coil, readily apparent in both FIG I and FIG 2 where valve spring means 3 contacts the coil valve spring base plate 14 at its plane interface. Note that the nozzle outer spiral fuel channel 24A and nozzle inner spiral fuel channel 25A both have three-sided cross sections comprised of the two circular adjacent coil sections and the linear edge of the cylindrical insertion tip bore 2 or the cylindrical center part, being either the axial fluid pipe 17 or the outer surface of the axial fluid channel coupling slide bearing 16.

At the axial level of the contact plane between the spring means 3 and its supporting coil valve spring base plate 14, wherein the two-dimensional shape of this contact was described as the crescent moon-shaped ground planar surface of the spiral coil spring, we find a practically unrestricted exit aperture for the three-sided nozzle outer spiral fuel channel 24A at the upper left-hand corner of the cross section of coil valve spring base plate 14 where this corner is near to the contact point of the circular cross section of the valve spring means 3, providing a direct and nearly immediate path to the adjacent inlet of a lower hydraulic chamber external channel 5A, and all other such channels surrounding the valve lift piston 9 by way of the circular wall of the insertion tip bore 2. And for the nozzle inner spiral fuel channel 25A at this same point we find fluid communication with the lower hydraulic chamber external channels 5A available via one of an unspecified number of valve closer inner flow hole I 5A, which depending on the outer diameter of coil valve spring base plate 14, and the size, location, orientation, and number of valve closer inner flow hole I 5A, may be inferior to the quality, in terms of flow cross section area, by comparison with the corresponding nozzle outer spiral fuel channel 24A communication.

Improvement of this more primitive flow-through design will be addressed by a technical solution in the embodiment of F1G 3 to come. Such a technical solution to be described later is not necessarily better in terms of minimized resistance to fluid flow and maximized fuel pressure at the nozzle holes 6 than the present simpler and less expensive solution of FIG 2, depending on complex viscous fluid flow. Therefore both solutions are offered as alternative choices to address specific fuel flow performance of any applicable embodiment.

Note that the crescent moon-shaped surface of mutual contact between the ground end surface of coil valve spring means 3 and spring base & piston channel unit 14A progressively blocks the coil valve spring fuel passages 15 as drawn, wherein there is no blockage on the left hand side, and complete blockage on the right hand side, with progressively increasing blockage of presumed but not shown radially disposed fuel passages 15 (in other planes containing the same nozzle valve axis line) from left to right in the view of the drawings. The valve closer spiral flow transfer annular cavity 26A (not present in FIG 2), introduced in the improved embodiment seen in FIG 3, improve this situation, especially in regard to flow rate exiting from the nozzle inner spiral fuel channel 25A.

At the valve closer inner flow hole ISA at the upper right hand corner of the cross section of coil valve spring base plate 14 we find apparent total blockage of fuel flow for both the nozzle outer spiral fuel channel 24A as well as for the nozzle inner spiral fuel channel 25k Such total blockage may be improved by limited grinding of the end of the spring, but that should be unnecessary given sufficient size and number and disposition of valve closer inner flow hole I 5A, including the valve closer spiral flow transfer annular cavity 26A to be introduced in the example embodiment of FIG 3, which in the case of the lower (downstream) support of the coil valve spring means 3 may be regarded as a constituent part (or sub-component) of the valve closer inner flow hole 1 5A.

Significant in FIG 2, by comparison with FIG 1, is to note that in FIG 2 the valve lift piston top 11 has come into contact with the cylinder top 12, the valve cylinder chamber 8 of FIG I having collapsed to nothing, and its incompressible fuel volume having been discharged into the low pressure valve closer unit fluid channel 13. Contact of valve lift piston top 11 with cylinder top 12 creates vacuum adhesion between the two smooth plane surfaces in contact submerged in a liquid. In other words, the valve lift piston top 11, by virtue of its flat, smooth plane surface tends to stick by vacuum adhesion (having expelled practically all liquid from between the two plane surfaces in contact immersed in ambient fluid) to the plane and smooth cylinder top 12 of the valve cylinder chamber 8, which to some extent inhibits whatever very small tendency there is for the ultra light valve closer unit 10 to rebound elastically from its collision, as discussed above.

Also noteworthy in FIG 2 is how the nozzle hole pressure equalizing groove 23 has risen from a position below the inlet orifices of the nozzle holes 6, where due to the VCO effect of the valve tip section 20 sealing the nozzle holes 6 sufficiently to prevent significant leakage of any of the trapped fuel below the pressure tight valve seating and below the inlet orifices of the nozzle holes 6, the fuel volume trapped within the nozzle hole pressure equalizing groove 23 rises very early in the valve lift trajectory (namely by a relatively short axial distance) to within low resistance (unimpeded) fluid communication with the nozzle holes 6, at which point of early nozzle flow the nozzle hole pressure equalizing groove 23 begins to act in an uniform manner around its circumference as a fuel collector at a relatively low flow rate per unit of cross sectional area at its circular perimeter with a relatively low viscous fluid flow resistance to inflow around its circular perimeter, and relatively high pressure due to that low flow rate and channels that fuel collected at a relatively high pressure to near proximity with the inlet orifices of the nozzle holes 6. Now what happens is that this relatively high pressure is presented immediately in front of the nozzle holes 6 at a flow rate which due to the relatively large cross sectoin of the nozzle hole pressure equalizing groove 23 is substantially greater than the uniformly distributed low inflow rate per cross sectional area into the nozzle hole pressure equalizing groove 23.

This discharge flow rate from the nozzle hole pressure equalizing groove 23 is significantly higher than in the absence of said groove 23 due to the relatively large cross section of the nozzle hole pressure equalizing groove 23 offering low resistance to hydraulic flow within its channel in comparison with the relatively small cross section area opening encircling the nozzle hole 6 orifice normally allowing passage of fuel into the orifice.

Furthermore, since the effect of nozzle hole pressure equalizing groove 23 is to create increased fuel pressure into the nozzle holes 6 at relatively lower valve lifts, the valve lift piston 9 travel limiting stop, which is cylinder top 12, may be positioned at a lower level of valve lift. The result is faster valve opening and closing times for an equivalent amount of fuel injection, because less time in travel is involved, and less mechanical wear and tear.

Whatever the cross sectional area of said groove 23, large or small, this channel augments, adds to, the available inflow cross section area surrounding each nozzle hole 6 orifice. There exists an optimum depth and width, or cross section area, of such a nozzle hole pressure equalizing groove 23 beyond which the nozzle hole 6 fuel pressure is not substantially increased, because the cross section area of the nozzle holes 6 themselves limit this flow at any given fuel rail pressure. A groove located at a distance of a fraction of the nozzle hole 6 diameter below the nozzle hole 6, which is about the same width as the nozzle hole 6, and half of its depth may possibly be near the maximum cross section limit of said groove 23 beyond which nozzle flow rate gains substantially diminish.

The fuel return flow system of the present invention avoids neccessity for an increaed capacity high pressure fuel pump to accommodate such return flow due to the fact that 100% of the returned flow to the fuel tank from the present fuel injector is due entirely to the extremely restricted flow-through of fuel past high precision and close-fitting sliding sleeve bearings. And that negligible quantity can only flow during the extremely brief intervals of the injection pulses, and is negligible if not zero during the nearly entire combustion cycle interval between the injection due to the fact that the piezoelectric metering valve unit (to be introduced later in the description for FIG 5, but mentioned out of order here, in advance, only to provide broader insight into aspects of FIG 2).

The tiny volume of fuel ejected from above the valve lift piston top 11 when it rises to open the nozzle valve, which is the volume of valve cylinder chamber 8 (seen in FIG 1), is always sucked back into the same volume from whence it was expelled by action of the valve spring means upon the closing nozzle valve, which forces the nozzle valve closed and at the same time sucks this tiny volume of fuel back into its previous volume, the valve cylinder chamber 8. And no flow restriction retards this operation. Therefore, in theory, the present system could operate at a zero net level of return fuel flow if there were no leakage losses past any of the slide bearings in the system, which is expected to be very nearly the case in practice.

The classic and well known nozzle valve opening force balance equation operational in the classic hydraulic fuel injector case, is that the net opening force upon the fuel injector valve at the instant of valve opening, just before the valve opens, is the spring force plus the hydraulic valve lift control chamber pressure times the the difference of the valve piston area minus the valve seat area. After the valve opens the fuel pressure upon the valve seat area drops out of the equation, increasing the valve opening force by that amount, which is generally regarded as beneficial for normal valve performance.

In the present invention, there are additional forces to be balanced in regard to the preceding calculation. One force of this additional balance is the force of ambient fuel rail pressure acting on the outside of the cavity of low pressure within the sliding axial fluid channel coupling slide bearing 16 slidingly engaged with the end of the axial fluid pipe 17 resulting in a force urging the two coupled sliding parts to slide together, which is a force acting to lift the nozzle valve from its valve seat.

And the second additional force unaccounted for in the classic equation is the fuel rail pressure within the insertion tip bore 2 acting on the shaft cross section area of the piston coupling and guide pipe 18 acting to close the nozzle valve. As can be seen in some of the variant example embodiments of the present inventive nozzle valve, such as that shown in FIG 3, the two circular cross section areas subject to equal and opposing pressures are close to being equal, substantially cancelling the effect of these additional forces.

In the example embodiments of FIG land FIG 2, for example, a piston coupling and guide pipe 18 diameter was increased, adding to it an annular shoulder for additional axial supportd of the pring base & coupler slide bearing unit I 4A for sake of illustrating non-limiting design options where necessary.

Drawing FIG 3, showing an example embodiment functionally identical to that of drawing FIG I, and mostly identical in structure, except for obvious structural and configurational differences, exhibits two differences from FIG I as improvements: a reduced diameter valve seat, and an alternative valve closer unit 10 assembly geometry, wherein the axial fluid channel coupling slide bearing 16, the coil valve spring base plate 14, and the piston coupling and guide pipe 18 are made of one piece as an integral unit termed the spring base & piston channel unit l4B, which is assembled having rotational freedom of relative movement within its smooth bearing bore in valve lift piston 9 to compose the valve closer unit 10. The present embodiment, besides increasing the outside wall thickness of the bottom of the tip end of the nozzle insertion tip I immediately above the nozzle holes 6 for improved fuel pressure resistance, incorporates the separate nozzle tip hydraulic cylinder group 4 into the nozzle insertion tip upper part IA as of one piece, providing options which are applicable, with variations, to all of the present invention embodiments.

Attachment of the nozzle insertion tip lower part I B to the nozzle insertion tip upper part IA may be done by either hot or cold metal joining processes, by brazing techniques, or by cold bonding using Surface Activated Bonding (SAB), accomplished for example by atomic surface activation of the surfaces to be joined by ion irradiation, fast atom beam (FAB) irradiation, or hydrogen radical beam irradiation. Other bonding variants include application of high pressure to the joint while heating the joint. A tooling cylindrical shaft may be left at the tip of nozzle insertion tip lower part I B for manipulating and centering the part during such processes, to be machined off later. Note that there are unlimited variations available for the division between nozzle insertion tip upper part IA and nozzle insertion tip lower part I B, wherein the simplest, and likely preferable for SAB bonding shows a plane surface joint, and the dotted lines being part of nozzle insertion tip lower part I B instead of the plane surface joint, represent a stepped joint possibly preferable with brazing.

The relevant discussions of FIG I and FIG 2 mostly apply to FIG 3 likewise, with the following differences: The reduced diameter of the valve seat of FIG 3 is suggested by the fact that the larger valve seat of FIG I is representative of a more or less typical 7mm VCO nozzle valve, (also the 9mm VCO nozzle) as shown in FIG 19. As explained in the discussion of the embodiment of FIG I above, the mass of the valve closer unit 10, which comprises the total dynamic mass of the element which closes the nozzle valve, is but a small fraction of the mass of a typical VCO (or sac) valve needle, which therefore impacts its valve seat with far less energy, producing less wear. This fact enables the valve seat contact area, and consequently its diameter to be reduced if desired, as can be seen by comparison of the valve tip congruent seating region 21 and the valve seat congruent seating region 22 of FIG I and FIG 3. The consequence of such diameter reduction is increased seating pressure of the valve in its seat (namely valve spring bias force per unit area of the valve seat), which improves sealing of the valve seat. There is a competition of advantages between better valve sealing through increased seating pressure and reduced valve seat wear through reduced seating pressure and kinetic impact. Reduced kinetic (momentum) impact of valve closer unit 10 on its valve seat in valve seat congruent seating region 22 will reduce valve seat wear, which enables reduction of the valve seat diameter if desired, for a given optimum valve spring means 3 bias force for the embodiment of FIG 1.

An advantage of reduced valve seating diameter is that the upper (upstream) edge of the valve seat and valve tip congruent seating regions 21 and 22 respectively may be moved significantly closer to the nozzle holes 6, which reduces viscous flow distance through the restricted gap (a problem of viscous flow in a duct) between said two regions 21 and 22 in order to reach the nozzle holes 6, thus increasing fuel pressure at said nozzle holes 6.

The actual "duct gap" herein referred to is revealed by comparison of FIG 2 with FIG 4, showing the space between said seating regions 21 and 22 having two different lengths.

Increasing the diameter of the piston coupling and guide pipe 18 may be done in a very similar, but in reversed direction manner as that shown in FIG 1, wherein the diameter of the piston coupling and guide pipe 18 penetrating the valve lift piston 9 remains as shown in FIG 3, but its diameter through the piston coupling and guide pipe bearing 19 is increased, with the resulting radial shoulder supported upon a radial surface of the valve lift piston 9.

The optional valve bounce elimination device for valve closer unit 10 mentioned in the description for FIG 1, flow restriction either by simply reducing the diameter of any part or all of valve closer unit fluid channel 13 to the degree that desired viscous damping of the motion of valve closer unit 10 is achieved (not shown), or by a more complicated flow restriction means disposed anywhere in valve closer unit fluid channel 13 (a deliberately leaking check valve, for instance, for uni-directional flow restriction, not shown), may be placed anywhere in rigid connection to valve closer unit 10 (or even outside of it), being advantageously within the valve closer unit fluid channel 13, which in the embodiment of FIG 3 includes piston coupling and guide pipe 18.

The present embodiment shows the improvement earlier promised for later presentation in the description of the embodiment of FIG 2 in regard to an improved flow capacity of the exit orifice of nozzle inner spiral fuel channel 25A, comprising an valve closer spiral flow transfer annular cavity 26A, which is in the present case merely an elaboration or extension (part and parcel) of the valve closer inner flow hole I 5A, which may be a groove, a hole, or a combination of both, so long as it provides fluid communication between at least the nozzle inner spiral fuel channel 25A and the lower hydraulic chamber external channels 5A. The advantage of such an valve closer spiral flow transfer annular cavity 26A and axial fluid pipe spiral flow transfer annular cavity 26B, at each end of the spiral valve spring means 3, (as well as a metering valve spiral flow transfer annular cavity 26C for a spring application for a pulsed fuel flow controlling (metering or dispensing) means to be first shown in FIG 5 and described for FIG 7), is that unlike discrete holes opening to the coil valve spring means supporting surface of the coil valve spring base plate 14, the annular length and radial width of the valve closer spiral flow transfer annular cavity 26A is uninterrupted, and may be of uniform width throughout its circumference. The only restriction to any number, disposition, or size of holes drilled through the coil valve spring base plate 14 for bringing the valve closer spiral flow transfer annular cavity 26A into fluid communication with the lower hydraulic chamber external channels 5A is preservation of the mechanical integrity and robustness of the coil valve spring base plate 14 in its function.

At the axial level of the contact plane between the spring means 3 and its supporting base plate 14, wherein the two-dimensional shape of this contact was described as the crescent moon-shaped ground planar surface of the spiral coil spring, we find a practically unrestricted exit aperture for the three-sided nozzle outer spiral fuel channel 24A at the upper right-hand corner of the cross section of coil valve spring base plate 14 where this corner is near to the contact point of the circular cross section of the valve spring means 3, providing a direct path to all of the inlets of lower hydraulic chamber external channels 5A, by way of the circular wall of the insertion tip bore 2. And for the nozzle inner spiral fuel channel 25A at this same point we find fluid communication with the valve closer spiral flow transfer annular cavity 26A, which in turn is in fluid communication with all of the unspecified number of valve closer inner flow hole ISA, all of which are in fluid communication with the lower hydraulic chamber external channels 5A.

Note that the crescent moon-shaped surface of mutual contact between the ground end surface of coil valve spring means 3 and coil valve spring base plate 14 would progressively block the inlet apertures of the coil valve spring fuel passages 15 as drawn if the valve closer spiral flow transfer annular cavity 26A were not present (which is the condition illustrated and described in this context for FIG 1), wherein there is no blockage on the right hand side, and complete blockage on the left hand side, with progressively increasing blockage of presumed but not shown radially disposed fuel passages 15 (in other planes containing the same nozzle valve axis line) from right to left in the view of FIG 3. The valve closer spiral flow transfer annular cavity 26A (not present in FIG 2) introduced in the improved embodiment seen in FIG 3 and onwards, may in some cases improve the overall flow-through rate in a small to moderate degree, depending on the number, size, and position of the valve closer inner flow hole ISA, especially in regard to flow rate exiting from the nozzle inner spiral fuel channel 25A.

The penalty in comparison with the simpler design of FIG I is slightly more complexity, and slightly greater thickness of the coil valve spring base plate 14. On the other hand, the advantages of improved fuel flow (reduced viscous flow resistance) offered by the present improvement may not always be needed. In other words, if the simpler design provides adequate performance, the improvement may not be needed. Therefore both options are provided.

It may be worth mentioning at this point is that exactly the same two spring seating options employing a valve closer spiral flow transfer annular cavity 26A in principle apply to the upper end of the spiral valve spring means 3 as well (and also in FIG 7), although the drawings do not show the obviously optional upper end analogue of the simpler of the two coil spring seating designs, analogous to that found in drawings FIG I and FIG 2, which avoids the valve closer spiral flow transfer annular cavity 26A.

Regarding precision centering and alignment and valve guidance to limit the problem of asymmetric injection spray patterns at low levels of valve needle lift, designers have stabilized the conventional VCO needle by adding a so-called double guided needle arrangement. The present invention should be immune to the problem of the needle tip wandering off of its axis, since the close proximity of the needle guidance to the valve seat (ie the nearest distance of the precision close fit bearing surface of the valve closer piston 9 in its cylinder to the valve seat, herein termed the valve closer unit guidance radial support point 27) would make such an off axis positioning of the valve tip congruent seating region 21 practically impossible. This stabilizing proximity may be conveniently represented by the diameter of the valve lift piston 9, wherein in the closed valve position, the axial distance from the lowest axial level of lateral (radial) valve guiding support for the valve closer unit 10 (ie the furthest such point in the downstream direction), to a point of contact between the valve tip congruent seating region 21 and its mating valve seat congruent seating region 22 (which is the target, or aim of said valve guiding radial support) is less than the diameter of the valve lift piston 9.

More specifically, an object of the present invention has been to limit as said the distance between a valve closer unit guidance radial support point 27 and its nearest point on a ring of valve sealing contact with the valve seat, which may lie anywhere between the circle of upstream edges of the inlet orifices of the nozzle holes 6 in the valve seat, and the upstream limit of "close proximity" between the valve tip congruent seating region 21 and the valve seat congruent seating region 22 when the valve is closed, which upstream limit specifically is the upstream edge of the valve seat congruent seating region 22.

As a possible specific definition of the meaning of "close proximity" of said congruent seating regions 21 and 22, here close proximity of the profiles (cross section) of the valve in its seat may be said to exist at the point on the valve seat profile where the tip of a 5.7° = arctangent 0.1 wedge touching the valve seat, can be inserted within an angle between the valve seat and its sealing valve tip. The two wedging sides of this wedge must be tangent to both sides of said gap at the points of wedge contact, and each of the sides must have length twice the width of the gap at the point where the wedge is tangent to the gap, without intersecting the valve cross section. Said "close proximity" ceases to exist at a line normal to the valve seat which passes through said tangent to the sealing valve tip.

According to this definition, the valve seat congruent seating region 22 and the valve tip congruent seating region 21 are in "close proximity" as drawn with the valve closed, except at the location of nozzle hole pressure equalizing groove 23, whose depth is 100 microns and whose width is 160 microns, and the nozzle holes 6 are 120 microns in diameter.

Drawing FIG 4 shows the embodiment of FIG 3 in the valve open position. As in comparison of FIG I with FIG 2, so also comparison of FIG 3 with FIG 4 shows distinctly the valve tip congruent seating region 21 and the valve seat congruent seating region 22.

The description of FIG 2 serves practically unchanged as the present description of FIG 4,taking into account their minor differences as presented in the description of FIG 3.

Drawing FIG 5 shows the nozzle valve embodiment similar to that of FIG 3 incorporated in an example embodiment of a complete VCO piezoelectric actuated fuel injector, wherein the acceleration and speed advantages of valve closer unit 10 should overcome the normal disadvantage of state of the art nozzle valve needles to bounce from impacts with a valve travel stop which due to the impossibility of controlling such bounce at the presently high valve needle speeds can not be used in the state of the art.

Items discussed in the present broad perspective view showing the entire fuel injector, such as the cylinder top 12 not shown in FIG 5 may be referred to in FIG 3 and FIG 4, as well as items possibly too small or fine to be clearly visible.

Little or no valve closer unit 10 bounce from impact with its cylinder top 12 travel stop in any of its embodiments is expected in the present invention due to its low inertial mass, low aspect ratio (ie low length compared to diameter), high hydraulic pressure opposing and resisting any possible low inertia bounce, the viscous hydraulic flow pressure gradient established between the conforming surfaces (in the present case plane surfaces) of the valve lift piston top 11 and the cylinder top 12 planes of contact of the moving open or closed valve closer unit 10 while the gap between said surfaces is much smaller than their diameter, which substantially retards their impact as well as any possible rebound following impact, familiarly known as the hydraulic adhesion on separation (or cushioning when approaching) effect of two planes in near proximity immersed in a viscous fluid medium (also responsible for the hydroplaning effect of automobile tyres on a wet road surface, and of the aerodynamic ground effect keeping an aeroplane aloft below normal flying speed while in near proximity to the ground), and likewise retards any residual bounce, and ambient fluid viscosity.

In addition to the foregoing pressure gradient effect is that, unlike plane surfaces immersed in a liquid, the edges or borders of the plane surfaces of the valve lift piston top hand its facing cylinder top 12 of the valve cylinder chamber 8 (see the enlarged view of valve closer unit 10 of FIG 3) are confined to and bounded by their valve cylinder chamber 8, which provides fluid communication only with the low pressure supply channel 36 and only through an aperture either in the valve lift piston top 11 via the valve closer unit fluid channel 13, or through the cylinder top 12 directly to the low pressure supply channel 36 (or an alternate low pressure source of a venturi) as will be seen in the embodiment of FIG 10, 11, 13 and 17. The smaller the orifice opening into said two end surfaces of valve cylinder chamber 8, the greater the magnitude of the hydraulic pressure gradient of viscous fluid flow, and its impact cushioning and adhesion based bounce retarding effects.

And if all else fails to remedy this bounce, the previously mentioned valve motion damping flow restrictor (not shown) or suitably reduced diameter disposed in the valve closer unit fluid channel 13 preferably (see FIG 1) of valve closer unit 10, or anywhere within the low pressure supply channel 36, the closer to the cylinder top 12 the better, will certainly remedy this bounce. These valve bounce controlling principles unique to the present invention, if only due to its crucial low inertial mass valve closer unit 10, may be combined and their parameters manipulated to perfect any design of the present invention. These principles apply equally to the embodiment of FIG 10, 11, 13 and 17.

The negligible inertial mass of the valve closer unit 10 being exceptionally responsive to valve opening and closing changes in control pressure of the fuel, which is the heart of the nozzle unit 28, responds to high-speed switched fuel rail pressure from the high pressure supply channel 34 as switched (or pulsed) by the piezoelectric actuated metering valve unit 29, and conducted through the switched flow supply channel 44, reaching the hydraulic valve lift control chamber 7 via connection with the cavity of insertion tip bore 2 in the nozzle insertion tip 1, where said switched fuel rail pressure acts to lift the valve closer unit 10. The switched fuel rail pressure is hydraulically unbalanced by atmospheric pressure supplied to the valve closer unit 10 through the axial fluid pipe 17.

Three connections are evident for the present embodiment: a fuel rail (or other high pressure fuel source) high pressure input pipe coupler 30 (left side); a low or atmospheric pressure fuel return channel connection 31 for return of fuel to the fuel tank (right side); and an electrical connector 32 (middle) for control of the piezoelectric actuator unit 33 of the piezoelectric metering valve unit 29. The high pressure input pipe coupler 30 (shown on FIG 5) connects to the high pressure supply channel 34, which leads only to the metering valve inlet chamber 35 of the cylindrical piezoelectric metering valve unit 29 (the crossing fuel channel is in a separate radial plane), which both actuates hydraulically as well as supplies with fuel the nozzle unit 28. Arrows in various fuel channels indicate flow direction when respective valves are open.

In order to provide a practically large hydraulic pressure differential for nozzle valve opening and closing operation of the nozzle unit 28 by piezoelectrically switched fuel rail pressure, a low pressure as low as atmospheric pressure connection is provided to the nozzle unit 28 from the fuel return channel connection 31 by way of the low pressure supply channel 36.

The piezoelectric stack element 37 of the piezoelectric actuator unit 33 expands in the axial direction upon energization in order to hydraulically lift the valve closure unit 10, causing fuel injection.

The complete VCO piezoelectric actuated fuel injector is assembled at its upper end as drawn (without further elaboration), by threaded attachment of the two fuel connectors and one electrical connector to a fuel injector body 38. And at its lower end, a threaded cap nut 39 retains the nozzle unit 28, attaching it to the fuel injector body 38 and its internal parts as illustrated.

Description of the embodiment of FIG 5 continues in its enlarged detail of FIG 6.

Drawings FIG 6 and FIG 6A show the fuel flow through the metering valve unit 29 and the nozzle unit 28 with its nozzle insertion tip 1, enlarging the view of the dynamic parts of the fuel injector of FIG 5, which are in the nozzle unit 28 as previously described in detail in drawings I through 6, and in the piezoelectric metering valve unit 29. The design of the piezoelectric metering valve unit 29 addresses the technical problem of overall compactness, and especially compactness in terms of axial length of the moving parts, as well as reduced inertial mass of the moving parts as a result of that compactness, enabling faster operating speed.

Drawing FIG 6A modifies FIG 6 to having a shorter valve spring means 3 length. A main purpose of this view, which shows how valve spring means 3 of different lengths are all equally well accommodated within the insertion tip bore 2, is to draw attention to the fact that any spring length may be adapted to the structure presented in both FIG 7 and FIG 8: The structure of the axial fluid pipe 17 above and below the interface with the top of the valve spring means 3, at the location of the axial fluid pipe spiral flow transfer annular cavity 26B, change in length but not in diameter when the axial fluid pipe spiral flow transfer annular cavity 26B is relocated from the axial level of the annular transfer chamber downstream into the insertion tip bore 2. The fuel injector body 38 is closed at its nozzle end by threaded cap nut 39.

In the present view, the axial fluid pipe 17 is seen to channel outflowing fuel to the low pressure supply channel 36 whose main purpose is to provide a differential pressure in conjunction with the high pressure of the fuel rail to open the valve lift piston 9 as previously described for FIG 1. At the top of the axial fluid pipe 17, surrounding its outlet orifice, is a recessed wrench socket (namely the recess apparent there) for tightening the threaded axial fluid pipe 17 into its seat during assembly of the nozzle unit 28.

Continuing the description from FIG 5, the high pressure input pipe coupler 30 (shown in FIG 5) connects to the high pressure supply channel 34, which leads to the metering valve inlet chamber 35 of the cylindrical piezoelectric metering valve unit 29, where the pressure is contained until the opening of mushroom valve 40 by electrical axial expansion of piezoelectric stack element 37 which transmits its downward (downstream) directed axial movement through piezoelectric pushrod unit 41 against a piezoelectric biasing spring means 42, which for the present example embodiment consists of a pair of specially configured series stacked Belleville springs. Fuel entering the open mushroom valve 40 and passing through the piezoelectric metering valve unit 29 (see description of FIG 7for a detailed discussion of the dynamic process) emerges from the metering valve outlet port 43 which feeds the switched flow supply channel 44 which supplies the annular transfer chamber 45 which supplies the insertion tip bore 2.

From the annular transfer chamber 45, fuel entering and passing through the insertion tip bore 2 flows through the nozzle outer spiral fuel channel 24A and the nozzle inner spiral fuel channel 25A defined respectively by the gaps between the spiral valve spring means 3 and the insertion tip bore 2, and the axial fluid pipe 17 (pIus the outer surface of the short axial fluid channel coupling slide bearing 16) as discussed for drawing FIG 1. However, in a manner analogous to that discussed for drawing FIG 1, the fuel flow entering the nozzle inner spiral fuel channel 25A is made less restricted by application of a axial fluid pipe spiral flow transfer annular cavity 26B, which in the present case is created by a short, calibrated length spring pressure adjustment tube 46 (for adjusting the compression force of the valve spring means 3), which has radially disposed (in a circle around the axis) and oriented axial fluid pipe inner flow hole I SB (where LP = Low Pressure, and analogous to the valve closer inner flow hole ISA of the coil valve spring base plate 14 of FIG 1) for feeding fuel into the axial fluid pipe spiral flow transfer annular cavity 26B which is formed by the gap between its inner cylindrical surface of the spring pressure adjustment tube 46 and the outer surface of the axial fluid pipe 17. In other words, the presence of spring pressure adjustment tube 46 surrounding the axial fluid pipe 17 creates a axial fluid pipe spiral flow transfer annular cavity 26B in form and function similar to valve closer spiral flow transfer annular cavity 26A in FIG 5, at the lower end of the spiral valve spring means 3.

This may often be a preferred alternative to simply supporting the upper end of the spiral valve spring means 3 against a plain annular shoulder whose axial position is adjustable by a spring pressure adjustment tube 46 without the radially disposed and oriented axial fluid pipe inner flow hole I SB (the simplest approach). Exactly the same working principle as described for lower valve closer spiral flow transfer annular cavity 26A applies to the upper axial fluid pipe spiral flow transfer annular cavity 26B of the present FIG 6, except that the inflow into the axial fluid pipe spiral flow transfer annular cavity 26B passes through radially oriented axial fluid pipe inner flow hole I SB, whereas outflow from its lower variant uses axially oriented valve closer inner flow hole ISA (and the third variant, the metering valve spiral flow transfer annular cavity 26C, to appear in FIG 7, uses no holes at all).

The design objective of metering valve unit 29 (preferably piezoelectric, but optionally solenoid operated) is to provide a robust, precise, and compact (having functionally optimum layout geometry) outward opening metering valve (exploiting fuel pressure to augment the closing spring force of the valve) having relatively low inertial mass and a large seat diameter in comparison to its unavoidably small direct piezoelectric actuated lift such as to provide sufficient valve flow-through area at the valve seat to insure minimal through the valve pressure drop while supplying necessary flow for both the nozzle holes 6 and the lifting of valve lift piston 9 to its pressure sealing position when in contact with cylinder top 12 (best seen in FIG 1). Low leakage to the low pressure supply channel 36 is also provided, by a telescoping valve guide bearing path for such leakage (piezoelectric metering valve unit 29 will be described in detail in FIG 7). Exceptionally high speed and high frequency operation is expected from a piezoelectric actuated metering valve unit 29.

Metering valve spiral flow transfer annular cavity 260 designed to relieve possible flow restrictions due to possible narrow coil to coil spacing of the metering valve spring 57 may be appreciated and understood better in the context of valve closer spiral flow transfer annular cavity 26A and axial fluid pipe spiral flow transfer annular cavity 26B, which perform identical functions in different places of the fuel injector.

Valve closer unit 10 is designed to be sensitively and quickly responsive to operation by the hydraulic pressure effect of switching on and off of metering valve unit 29 according to features already described in preceding drawings, when piezoelectrically actuated and connected to a high pressure fuel rail as its pressure source.

Drawing FIG 7 and FIG 7A show in two valve diameter variants an enlarged view of the essential parts of the piezoelectric metering valve unit 29 (seen in its entirety in FIG 5 and FIG 6) in the closed position, and its interface with the bottom part of the piezoelectric actuator unit 32 (cf FIG 5) which drives it. The virtues of this design are compactness in both axial length and diameter, relatively light weight of moving parts for both high speed operation and low valve seat wear, precision not likely to deteriorate over a long service life, robustness to handle extreme fuel pressures, low leakage through its sliding bearings while open only, negligible leakage while closed, and high performance in terms of both speed of operation as well as flow-through rates to satisfy demanding fuel injector applications.

Continuing the description from FIG 6 to the present enlargement, the high pressure supply channel 34, which leads to the metering valve inlet chamber 35 of the piezoelectric metering valve unit 29 (cf FIG 6 for fuller perspective), where the pressure is contained until the opening of mushroom valve 40 by electrical energization axial expansion of piezoelectric stack element 37 which transmits its downward (downstream) directed axial movement through piezoelectric pushrod unit 41 against a piezoelectric biasing spring means 42, which for the present example embodiment consists of a pair of specially configured series stacked Belleville springs. The purpose of these piezoelectric biasing spring means 42 is to provide necessary compressive bias force against the piezoelectric stack element 37 for its internal operation, and to dynamically support its moving lower end, allowing axial movement, but no radial movement for the piezoelectric actuator unit 33 (can only be entirely seen in FIG 5).

What is special about these BelIevilIe springs is that they avoid, or at least reduce to a minimum, frictional contact. Their series stacked contact at their central hole areas, avoids contact with the shaft passing through these holes. Their nonfrictional contact with their surrounding mechanism is at only two points: at the circular contact with the piezoelectric push rod unit 41 where the expanding outer diameter of the upper Belleville spring at its bias, or working force, is calculated to closely match the inner diameter of a shallow cylindrical retaining recess (or raised annular retaining shoulder) in the BeIlevilIe spring supporting face of the piezoelectric pushrod unit 41, such that said Belleville spring is captured, or held radially immobile by the shoulder of said shallow cylindrical retaining recess. Similarly, the lower Belleville spring is calculated to expand to a radially immobile position in circular contact with its containing bore (cf FIG 6) at its bias, or working force.

The stroke length of the piezoelectric actuator unit 33 (cf FIG 5) is very small in proportion to its diameter, such that any movement with frictional effect at the supporting surfaces of the Belleville springs will be negligible. Also preferably the Belleville springs should be sufficiently thin such that their outer rims will be flattened to some extent at their bias force, avoiding the fatigue limit of internal Belleville spring stress, further resisting abrasive frictional supporting contact due to radial expansion of the spring. The contact between the two series stacked Belleville springs may take a variety of forms. The simplest is simple edge contact between identical diameter center holes of the springs, wherein these are held radially immobile by capture of each spring in its respective cylindrical recess.

Another solution is to make one hole smaller than the other, and depend upon the smaller tipped frustoconical Belleville spring being radially held by its surrounding hole of the opposing frustoconical spring in which it is seated. This eliminates the need for a recess to capture (radially immobilize) the upper spring. A third option is that which is illustrated, wherein the lower, larger spring has a smaller central hole than the upper spring (or alternately, vice-versa), but in addition this hole has been bent to form an annular central shoulder (flange) to radially immobilize the upper spring in its central hole, which again eliminates the need for a shoulder to prevent radial movement at its large upper contact edge. Such variants eliminate contact and friction of the Belleville springs with the central shaft of the piezoelectric pushrod unit 41, as well as any significant friction at their large diameter supporting edges.

Upon energization of the piezoelectric actuator unit 33, which overcomes the bias pressure of its piezoelectric biasing spring means 42, the central pushrod of piezoelectric pushrod unit 41 moves forward through a very small shaft gap 47 between its flat end face and the flat end face of the upper shaft of mushroom valve 40. A pressure relief channel to accommodate the collapse of the shaft gap 47 fluid volume may or may not be needed (is not depicted), depending on the bearing clearance of its two adjacent shaft bearings. The purpose of this small shaft gap 47 is to compensate for valve seat wear, wherein the mushroom valve 40 sinks upward into its valve seat, gradually reducing said gap, and also manufacturing tolerances, and temperature effects. It insures against the catastrophic possibility of the mushroom valve 40 being held open so as to leak. Voltage amplitude control of the piezoelectric stack could make fine stroke length adjustments to further compensate for such effects, as well as timing control.

Conventionally in fuel injectors, a hydraulic coupler is disposed between the piezoelectric stack and the valve it operates to eliminate all gaps and irregularities as mentioned, which could easily replace the shaft gap 47 of the present invention if necessary, but this may not be necessary, especially because the operation of the nozzle valve is not sensitive to the precision of the maximum valve gap of the metering valve unit 29, provided that that gap is provided an excess (ie margin or buffer of error) of valve lift which includes anticipated gap irregularities, due to the flow limiting, throttling effect of the nozzle holes 6, in combination with the fixed valve lift of the valve closer unit 10 due to its fixed valve stop cylinder top 12.

With these provisions insuring regularity of operation, the hydraulic coupler may not be necessary, and all that is necessary is to insure that this gap exceeds a certain operational minimum, but not by a large margin, since excessive valve lift may be detrimental to valve closure speed, as will be discussed later.

Thus, the primary design objective of the piezoelectric metering valve unit 29 was to provide a large inward-opening valve seat flow-through area at the typical relatively low 0.2% of piezoelectric stack length lift in order to meet the nozzle flow demand.

Such a relatively large valve seat in turn needs a correspondingly increased precision of centering and alignment of the piezoelectric metering valve unit 29. Because small eccentricities will have a correspondingly larger effect on leakage at valve closure for a low-lift, large diameter valve seat than for a small diameter, high lift equivalent flow poppet valve. Small diameter metering control valves depend on their relatively large valve lift to produce an equivalent flow-through, and need not provide as close valve guiding shaft tolerances and precision for centering of their relatively small diameter valve in its seat.

The same eccentricity (radial imprecision) which causes a small leak in a small diameter poppet valve will cause a correspondingly larger leak in a large diameter poppet valve.

An effect favoring assured seating of eccentrically closing large diameter poppet valves over small diameter ones is the high pressure effect forcing the larger closed valve with overwhelming force into its seat, unless a manufacturing defect of an eccentric bearing prevents the valve from closing completely. In other words, what human precision is incapable of achieving in order to precisely center the large diameter poppet valve in its seat, the high pressure at which the valve works forces to happen, given sufficient radial play. For a small diameter poppet valve, such seating force and effect due to the valve working pressure is correspondingly smaller, For outward opening poppet valves operating against great pressure, a larger valve seat diameter is equivalent to having a stronger valve spring, and therefore, a better valve centering and sealing effect due to the closing pressure.

For these reasons, the present design of piezoelectric metering valve unit 29 provides an advantageously compact telescoping triple sliding bearing guidance of the double stem metering mushroom valve 40 for precisely centering the valve which must be made exceptionally robust and precise due to its large seat diameter in combination with its high working pressure and high valve opening and closing forces.

The conveniently implemented outer stem sliding bearing 48 of mushroom valve 40, disposed in the center of the metering valve inlet chamber 35 floor", may be configured to be positioned axially as close to the valve seat as needed for optimum valve centering precision. To prevent fluid compression lock in the blind bearing hole of the outer stem sliding bearing 48, a blind hole pressure relief channel 49 creating fluid communication between the trapped volume in said blind bearing hole and the metering valve inlet chamber 35 is bored into the valve stem, which may pass through much of its length to lighten the valve if desired, wherein said blind hole pressure relief channel 49 comprises at least one intersecting hole is bored to provide fluid communication between the trapped volume at the bottom of the bearing bore and the metering valve inlet chamber 35. Of course, flutes or grooves in the surface of the outer stem sliding bearing 48 could perform the same pressure relief function.

The outlet side stem of mushroom valve 40 is guided by two concentric slide bearings. The outer of these concentric bearings is the larger of the two concentric bores of recessed bearing valve spring retainer 50. Its smaller concentric bore is close fitted to the outlet stem of the mushroom valve 40 and locked in place with a split valve lock 51 (in the present example conveniently like a normal automobile cylinder valve spring retainer lock).

Said outer concentric bore of the recessed bearing valve spring retainer 50 may compactly engage by sliding bearing contact the outer bearing surface of the central inwardly projecting cylindrical part of the coaxial cylinder spring housing 52, which is its bored coaxial inward-projecting cylinder 53 comprising one or more inner bore diameters hosting two sliding bearing shafts: the shaft of piezoelectric push rod unit 41, and the outlet stem of mushroom valve 40, with the shaft gap 47 in between. This latter nested bearing arrangement, in combination with the particularly robust and precise outer stem sliding bearing 48 arrangement results in a particularly compact, robust, and effective low inertial mass valve geometry suitable for precision high pressure, high-speed valve applications.

All of these bearings are expected to be of very close tolerances due to the miniature size of the valve mechanism and its requirement for high precision and demanding performance, with the result that the high pressure differential (expected full fuel rail pressure) which exists longitudinally through the bearing clearance of the central bore of the bored coaxial inward-projecting cylinder 53 substantially restricts high pressure fuel leakage into the cavity housing the piezoelectric biasing spring means 42. This expected small leakage flows out of said cavity via the metering valve leakage port 54 connecting immediately with the low pressure supply channel 36 which conducts out of the fuel injector, the fuel tank being the final destination of the fuel. Hydraulic lock relief channels relieve telescoping movement pressure changes in the trapped volume at the bottom of the coaxial cylinder spring housing 52. Said holes 50 may, of course, be substituted by any kind of fuel passage to the ambient fuel volume, typically a groove or a hole, or a hole enlargement so as to entirely eliminate a bearing surface. In addition, Hydraulic lock relief channels 55 may serve as a viscous damping for valve closure action if desired, by restricting their cross sectional area, to function as viscous flow restrictors, although such a function is not of particular interest for the present invention..

Note that the state of the art has perfected means of sealing against high pressure fuel leakage past the sliding bearing clearances of the metering valve unit 29 so as to affect adversely the piezoelectric actuator unit 33 (shown in FIG 5), which in the present embodiment has been conveniently and simply solved by the metering valve leakage port 54, which is obviously a solution replaceable by such means which eliminate back flow to the fuel tank entirely, where needed or desired, such as, for example, in the embodiments of FIG 15 and FIG 17.

Notable compactness with maximum precision of the metering valve unit 29 is achieved by close-fitting coaxial (one inside the other) assembly of its two cylindrical valve housing components, one inside the other, assuring maximum radial and axial precision of assembly. The outer diameter of the cylindrical coaxial cylinder spring housing 52 is sized for coaxial tight-fitting assembly with the inner diameter of the metering valve seat cylinder 56 as shown.

The assembly is accomplished by first assembling the mushroom valve 40 into its valve seat in the metering valve seat cylinder 56. Next the metering valve spring 57 is assembled to the preceding assembly. Next the recessed bearing valve spring retainer 50 is installed upon the outlet stem of mushroom valve 40, first compressing the metering valve spring 57, and then locking the recessed bearing valve spring retainer 50 in place with split valve lock 51. Finally, the coaxial cylinder spring housing 52 is inserted within the inner bore of the metering valve seat cylinder 56 of the thus far assembled unit, completing the metering valve unit 29 assembly with the exception of the relatively simple attachment of the housing for the metering valve inlet chamber 35, which in the present design takes place when both of these items are assembled by close-fitting insertion into a bore of the fuel injector body 38.

It is to be noted in the preceding assembly, that the metering valve spiral flow transfer annular cavity 260 which functions as a collector of fuel from the metering valve outer spiral fuel channel 24B, for channeling it to the metering valve outlet port 43 is formed by cutting the length of the skirt (the outer cylindrical part) of the recessed bearing valve spring retainer 50 cylinder short by an axial distance such that the metering valve spiral flow transfer annular cavity 26C is precisely formed as the resulting gap of said axial distance when the two cylinders are assembled. In other words, if the outer surface of the skirt of the recessed bearing valve spring retainer 50 cylinder were to be in full axial length contact with inner surface of the metering valve seat cylinder 56, there would be no metering valve spiral flow transfer annular cavity 26C. Therefore, to create the metering valve spiral flow transfer annular cavity 26C within the metering valve seat cylinder 56, one must cut short the skirt of the recessed bearing valve spring retainer 50 cylinder by the axial height of the metering valve spiral flow transfer annular cavity 26C as shown in FIG 7.

It should be noted that the piezoelectric metering valve unit 29 is not dependent on a low pressure supply channel 36 for its normal operation. For example, the metering valve leakage port 54 could have been eliminated by using a flexible high pressure seal to isolate the piezoelectric stack element 37 from the high pressure fuel supply, such as are commonly used in fuel injectors for protection of the piezoelectric element.

The cylindrical design of piezoelectric metering valve unit 29 is robust, providing precision valve seating throughout a long working lifetime. It is notably compact in both diameter as well as height, the cylindrical cross section being nearly square.

Its central inflow at the valve seat of mushroom valve 40 is disposed at the shortest possible distance to its metering valve outlet port 43 in its side, yet there is a flow path between the gap in the coils of the metering valve spring 57 along its entire length and circumference, where there exist metering valve outer spiral fuel channel 24B and metering valve inner spiral fuel channel 25B in form and function like the nozzle outer spiral fuel channel 24A and the nozzle inner spiral fuel channel 25A of valve spring means 3 of the nozzle valve (cf FIG 1), as previously discussed. Thus fuel flow follows a spiral path on the inside of the coil upward along the metering valve inner spiral fuel channel 25B analogous to nozzle inner spiral fuel channel 25A in the nozzle valve as described for FIG 2, where this fuel flow passes radially through the long spiral gap between the coils of the metering valve spring 57, and returns downward along the metering valve outer spiral fuel channel 24B analogous to the previously discussed nozzle outer spiral fuel channel 24A of the nozzle valve until it reaches the the metering valve spiral flow transfer annular cavity 26C, analogous in form and function to the valve closer spiral flow transfer annular cavity 26A of the nozzle valve, emptying into the metering valve outlet port 43 of the piezoelectric metering valve unit 29, which feeds the switched flow supply channel 44 to the nozzle valve. These spiral fuel channels distribute radial fuel flow-through throughout a long but possibly narrow spiral gap between the coils of the metering valve spring 57. This radial flow-through may, if needed, be augmented by axial flow along gaps between the outer diameter of the metering valve spring 57 and its adjacent cylindrical wall, and the similar gap between the inner diameter of the metering valve spring 57 and its adjacent cylindrical wall. Such fuel flow gaps, while permissibly quite small if not completely negligible in comparison with the total cross sectional area of all of the nozzle holes 6 (cf FIG 1), should be considered for size to ensure adequate fuel flow-through depending on free-standing (unguided) spring height (which like for typical intake and exhaust valve springs have a limited free-standing length to diameter aspect ratio) to avoid undue pressure drop through the valve besides reducing spring friction with the adjacent inner and outer cylindrical walls.

A noteworthy principle hidden in plain view in FIG 7 is that the relatively large mushroom valve 40 diameter of the piezoelectric metering valve unit 29 is combined with a relatively short, and therefore a relatively short-stroke piezoelectric actuator unit 33 producing a relatively low mushroom valve 40 lift. The implication of this is that a smaller mushroom valve 40 diameter piezoelectric metering valve unit 29 would require a greater mushroom valve 40 lift from a longer piezoelectric actuator unit 33 (having a longer piezoelectric stack element 37) to produce the same flow rate. The complete fuel injector as drawn in FIG 5 has unusually short overall length for a modern fuel injector. What this means is if a smaller diameter mushroom valve 40 were needed in order to open against a much larger fuel pressure (eg. fuel rail pressure), such a smaller diameter mushroom valve 40 will be easily accommodated by use of a longer piezoelectric actuator unit 33 for a longer working stroke, and which can also be made larger in diameter so that the longer working stroke will also be more powerful.

The valve seat diameter of the mushroom valve 40 is relatively much larger than its valve lift, so as to provide a relatively large opening fluid inflow cross section through the narrow opening at the valve seat, enough to both maintain high, nearly fuel rail pressure in the nozzle valve (cf FIG 1) at the nozzle holes 6 of the fuel injector (whose diameter should be small enough to limit the fuel flow) while lifting and holding the valve lift piston 9 (cf FIG 1) against its travel stop, the valve lift piston top 11 (cf FIG 1). When the valve lift piston 9 contacts its stopping valve lift piston top 11, fuel flow through valve closer unit fluid channel 13 (cf FIG 1) is blocked and shutoff, with minimal leakage of fuel through the sliding bearings of the valve lifting mechanism. Therefore, the extremely fast piezoelectric actuated full opening of mushroom valve 40 for fuel flow and pressure almost instantaneously lifts the nozzle valve valve closer unit 10 (cf FIG 1) to its travel stop, producing practically full fuel rail pressure at the nozzle holes 6, permitting fuel pressure to be applied far in excess of the valve spring means 3 force, producing fast opening and closing valve operation (in contrast to the state of the art, which can not benefit by the advantage of a travel stop for both increased speed of operation and higher nozzle pressures without regard to valve spring force).

Also, it takes much less force to keep the mushroom valve 40 open once it has been opened, than the original force required to lift it off of its valve seat, once pressure equilibrium between the metering valve inlet chamber 35 and the interior of the metering valve unit 29 has been almost instantaneously attained due to the characteristically enormous force and speed of the piezoelectric actuator unit 33. The fast attainment of this pressure equilibrium is assisted by the relatively much smaller cross section area of the nozzle holes 6 (cf FIG 1), limiting the flow from the metering valve unit 29. At full pressure equiiibrium, the piezoelectric stack element 37 is completely relieved of the fuel rail force previously acting on the mushroom valve 40 when it was closed, and is subject only to the metering valve spring 57 force, which could be many times smaller than the force due to the fuel rail pressure. The result of this phenomenon is that the force of the piezoelectric stack element 37 will lift the mushroom valve 40 considerably higher than the opening force would suggest, and this valve lift opens with an acceleration effect.

The closing speed of the mushroom valve 40 is likewise accelerated by its most likely exponentially rising pressure gradient due to the positive feedback effect whereby said gradient increases as the valve gap distance (termed,,throttling') closes. When the diminishing cross section area of the closing mushroom valve annular opening gap approaches the cross section area of the nozzle valves 6, a pressure gradient is established across said gap, which will create more closing force on the the valve than the metering valve spring 57.

These metering valve unit 29 opening and closing acceleration effects (suggesting a snapping open and a snapping shut effect of said valve due to the high fuel rail pressure) suggests a bonus effect of the outward opening mushroom valve metering valve unit 29 design which would speed up the opening time and the closing time of the nozzle valve, whose uniquely low inertial mass is intended to react sensitively and quickly to said snapping metering valve unit 29 action which is a high priority objective of the present invention.

Therefore, the critical requirement of the piezoelectric stack element 37 is that it must be able to lift the mushroom valve 40 against the fuel rail force on the mushroom valve seat area by a distance which relieves the nozzle holes 6 flow rate limited pressure gradient just described. Once that lift is attained, then the piezoelectric stack element 37, relieved of enormous fuel pressure force, could possibly double its stroke to that point, which is important for producing sufficient flow to quickly lift the valve lift piston 9 of the valve closer unit 10 (see FIG 1), involving only the very small total flow volume of valve cylinder chamber 8.

Another advantage for the present invention in all of its embodiments (cf all of the embodiment drawings), is that the sharp (very short duration) snap open and snap closed (or pop open and pop closed) character of the outward opening metering valve unit 29 due to the very high pressure gradient established across its valve seat at the low valve lifts where a substantial part of its operation takes place (due to the characteristically low stroke of any piezoelectric actuator unit 33 (shown only in FIG 5), creates especially abrupt (sharply rising and falling) opening and closing of the metering valve unit 29, and in turn of the valve closer unit 10 of the nozzie valve, which enhances its high speed and high frequency operation. In other words, an advantage of the present invention exploits what may otherwise be regarded as a weakness or liability of piezoelectric valve actuation, namely its limitation to produce very small operational stroke and valve lift. Herein, the flow capacity of the metering valve is increased not principally by the limited piezoelectric valve lift, but by its valve seat diameter, the opposition of whose enormous hydraulic pressure determined opening force is countered by the enormous force capability of the piezoelectric actuator, as balanced within practical limits.

In the process of that lift, once valve flow approaches the throttling flow-through limit of the nozzle holes 6, the pressure gradient through the mushroom valve 40 valve seat quickly disappears along with the valve closing force of that pressure gradient, leaving only the enormous piezoelectric force to snap the valve open. The net result of this process is that it suffices for the piezoelectric actuator to open the valve merely by a small amount as reduced to a small level by the enormous opposing hydraulic valve closing force, which nullifies this hydraulic closing force, suddenly, as by a "pop" or a "snap", enabling the full no-load stroke of the piezoelectric actuator to be achieved. And that valve lift, due to the low stroke nature of the piezoelectric operational principle, typically does not to go far beyond the point where the pressure gradient developed by fuel flow through the metering valve becomes substantial enough to cause a sharp (quick) "snap", or a "pop" of sudden valve opening by the immense piezoelectric speed and force working against the pressure gradient force through the opening valve, or the same pressure gradient closing the valve by a positive feedback effect accelerating the valve shut, which is a significant advantage for the speed and precision of the metering valve operation, but more importantly, when coupled with the use of a valve closer unit 10 of exceptionally low inertial mass, by virtue of which it is highly sensitive and responsive to the influence of hydraulic pressure upon it, this quick change in hydraulic pressure is translated into a quick change in the velocity of valve closer unit 10 to either open or close, resulting in substantial improvement of valve speed and frequency of operation over the state of the art, historically burdened by excessive nozzle valve inertial mass.

Moreover, when the metering valve is fully piezoelectrically opened, the valve closing force of its valve spring is maximum. And as this metering valve closes, instead of the closing force upon the valve diminishing, due to the reducing compression of its valve spring, the closing force on the metering valve increases due to the increasing pressure gradient across the valve seat boundary (inside to outside the valve). This rapidly accelerating hydraulic pressure gradient effect established by a positive feedback condition substantially accelerates the valve closure, wherein the piezoelectric mechanism does not puli the vaive shut by a rigid contact as it pushes it when opening the vaive, wherein the high speed of the piezoelectric principle accelerates the valve opening much faster than any mechanical valve opening alternative. It is in closing the valve that there is no dependence on mechanical coupling with this mechanism, other than a possible retarding influence. And it is just at the closing of the metering valve that the presence of this "snap" or "pop" valve closing principle insures that the metering valve will close much faster than in its absence, and probably keeping pace with the piezoelectric speed due to the enormous force of the fuel pressure upon the closing metering valve due to said valve seat pressure gradient, which is conditioned by the flow-through rate of the nozzle holes 6, and reduced by the pressure gradient through the nozzle holes 6.

The technical effect of this peculiar "snap" action of the outward opening, mushroom valve type metering valve unit 29 is inventively exploited in combination with the uniquely low inertial mass valve closer unit 10 of all of the embodiments of the present invention, which due to that low inertial mass are uniquely sensitive and responsive to change of applied valve actuating hydraulic pressure, to greatly exceed the previous performance limits of the state of the art for direct injection inwardly opening differential valve type nozzle" based fuel injectors, otherwise characterized as operating in a broad sense (allowing for inventive improvement) by the principle of,,nozzle differential ratio", often generically termed 25,,hydraulic" (also known as a,,mechanical").

Flexibility of configuration of the piezoelectric metering valve unit 29 is demonstrated in FIG 12, showing an alternative option of routing fuel through through the valve. This demonstrates how fuel can be routed both axially as well as radially from the metering valve spiral flow transfer annular cavity 26C all around the circumference of the piezoelectric metering valve unit 29. In other words, metering valve outlet port 43 may be located in a wide variety of places to suit the designer.

Drawing FIG 8A S and FIG SB 8A show in two valve diameter variants the piezoelectric metering valve unit 29 in the open position. Mention of fuel flow through the valve in the description for FIG 7 showing the valve in the closed position apply to the valve open illustration of FIG 8. Again, in order to provide an adequate fuel flow-through for the small valve lift of 0.2% of piezoelectric stack length, the present design applies a correspondingly large valve diameter as a simple solution, using a robust design capable of high precision operation required by the relatively large 4.0mm (4000pm) valve diameter. The valve lift depicted for illustration purposes only in FIG 8 (based on a 7mm nozzle insertion tip I diameter) is 87pm.

A common piezoelectric actuator stack for fuel injectors is 30mm in length, and produces a lift of 4Opm. So the piezoelectric stack would need to be about 70mm in length for the lift shown against zero force of fuel pressure, whereas the piezoelectric stack element 37 as drawn in FIG 5 is about 36mm in length. However, doubling this length would pose no practical problem, if a 9Opm valve lift were needed. The gap area at 0.09mm lift as drawn is 1.131mm2. The diameter of a nozzle hole 6 as drawn is 120pm, which is a fairly normal size at present for a 7mm nozzle size. This nozzle has an area of 0.0113mm2. A maximum limit for the number of such nozzle holes may be around 12, which would produce a flow-through area of 12 x 0.0113mm2 = 0.136mm2. Therefore, the piezoelectric metering valve unit 29, in the proportions in which it is drawn, has 1.131/0.136 = 8.3 times the flow-through area of 12 normal nozzle holes. Half of that valve diameter, or 2mm, would still give 4.2 times the minimum flow at only 1/4 of the force required to open the valve. Six nozzle holes (or half of the exaggerated number used for insuring a safe margin of error to produce the present drawings) is a normal number. Therefore it is safe to say that the depicted 9Opm valve lift is about double of what should guarantee an effective valve lift.

Or alternatively to produce half of the depicted flow through capacity of the metering valve unit 29, the 4mm diameter of the mushroom valve 40 (see FIG 7 for the reference number) may be reduced by half to 2mm, which would reduce the force of pressure on the valve by a factor of 4, which may be a preferred choice for some piezoelectric actuators.

Calculating the force needed by the piezoelectric stack to open the depicted 4mm diameter mushroom valve 40, we find a typical piezoelectric stack which produces 30pm lift at zero force having a stiffness of 200N/pm produces a maximum force at zero stroke of 6000 N, or about 600kg. Assume its length being similar to the 30mm length mentioned above producing 4Opm stroke. The force produced against the 4mm mushroom valve 40 at 2000 bar pressure is 251 kg. (At 2000 bar, a 2mm diameter valve producing 4.2 times the minimum required flow at the depicted valve opening would require 1/4 x 250kg = 63kg opening force, being more reasonable for the proportions and dimensions actually depicted in FIG 8.) Therefore the lift of this piezoelectric stack against 250kg force will be 250/600 x 3Opm = 12.Spm. Since 9Opm lift produces 8.3 times the flow-through area of 12 120pm nozzle holes, 12.Spm lift will produce 1.15 times the flow-through area of 12 120pm nozzle holes, which is double the lift needed to equal the flow-through area of a more normal number of 6 nozzle holes. In other words, a lift of 12.5pm is a reasonable approximation for the lift needed to serve 6 nozzle holes against a pressure of 2000 bar, and there is no problem to increase both the lift distance and the force by several times using larger piezoelectric stacks which would easily be accommodated within a normal sized injector body having a 7mm nozzle.

The above calculation which provides guaranteed minimum piezoelectric stack specifications for opening the mushroom valve 40 did not take into account the fact that when the pressure outside and inside the piezoelectric metering valve unit 29 equalize when the valve is open, then the immense force of the fuel rail pressure (2000 bar producing 250kg of force against the valve) suddenly disappears, leaving only the valve spring pressure, which is relatively negligible. Therefore, practically the full zero force stroke of the piezoelectric element will be realized in practice so long as the valve gap is equal to or larger than the total nozzle valve flow-through area, and the action of the mushroom valve 40 will be to snap open in about 100 microseconds due to the speed of the piezoelectric force applied, wherein the fuel pressure force on the opening valve falls due to the fall of the viscous hydraulic flow pressure gradient through the opening valve until it disappears when the nozzle hole area becomes smaller than the valve gap area, at a piezoelectric speed of about 100 microseconds; and to snap shut at about the same speed due to the fuel rail pressure acting on the closing valve when the piezoelectric force is removed causing a pressure gradient to arise through the closing valve gap which has a smaller cross section area than the area of the nozzle holes, creating very short (sharp) fuel pressure rise and fall times ideally suited for actuating the low inertial mass, and therefore highly pressure responsive and negligible bounce (easily suppressed bounce) embodiments of the valve closer unit 10 of the present invention.

It should be a design consideration to limit the mushroom valve 40 lift to not substantially exceed the limits of the pressure gradient effect through the opening and closing valve, which pressure gradient disappears when the throttling of the nozzle holes 6 limit the flow through the metering valve unit 29, because its valve spring may possibly not be able to keep pace with the 100 microseconds fall time of the piezoelectric stack, but the fuel rail pressure acting by means of the accelerating pressure gradient through the valve seat will have that ability, insuring the high speed of valve closure.

Another factor to consider is that during the approximately 95% of the combustion cycle when the fuel injector is idle there will be a possible limited bleed off of the residual valve closing fuel pressure through the sliding bearings of the valve closer unit 10. However, the current state of the art is able to execute as many as 7 fuel injections per combustion event, during the 5% of the cycle remaining. The present invention should be able to meet or exceed such performance, and between such rapid injections there will be no drop in the residual fuel pressure after the valve is closed. What this means is that the pressure opposing the piezoelectric stack between the rapid injections will be significantly less than the full fuel rail pressure, in the present example, 2000 bar, which will considerably relieve the performance demand on the piezoelectric stack and related metal components for most of the fuel injector service life.

Drawing FIG 9 represents a cross section of the hydraulic cylinder portion of the nozzle tip hydraulic cylinder group 4, wherein is illustrated the option of using drilled axially oriented holes for the lower hydraulic chamber external channels 5A, and is immediately related to drawing FIG I which is the broader context of the present discussion, and demonstrates by its relative proportions, and its use of a roughly estimated maximum number (maximum density) of lower hydraulic chamber external channels 5A of maximum estimated proportional diameter the ability of the lower hydraulic chamber external channels 5A to withstand the high pressures of modern fuel rails. The steel wall thickness of said channels 5A in FIG 9 are comparable in terms of ratio of channel diameter to wall thickness to that found in the insertion tip bores of modern direct injection Diesel fuel injectors operating at modern fuel rail pressures.

Drawing FIG 9 shows, in the context of the ability of fuel channels to withstand modern fuel rail pressures, the upper limits of the diameter lower hydraulic chamber external channel 5A relative to the insertion tip bore 2 diameter which is identical to the nozzle tip hydraulic cylinder group 4 outer diameter. The fuel carrying capacity of the structure of FIG 9 represents 12 times the fuel flow-through of 12 100pm nozzle holes, an unusually high flow-through rate, wherein the diameter of the lower hydraulic chamber external channels 5A could be reduced to 113 of their diameter shown in FIG 9 without affecting the performance of such a fuel injector, having 12 100pm nozzle holes.

The next important major item below the valve spring means 3 In FIG I is the nozzle tip hydraulic cylinder group 4, which is rigidly joined to insertion tip bore 2 (typically 4mm in diameter in a 7mm nozzle insertion tip 1), which rigid joint must be sufficient to withstand the constant impacts (collisions) of the nozzle valve parts described below tending to drive the separate unit version of the nozzle tip hydraulic cylinder group 4 upwards (upstream) in the insertion tip bore 2, as well as the cyclic swelling and expanding elastic response (on the order of a micron or more) of the steel nozzle insertion tip I to the fuel rail pressure injection pulses as a balloon expanding to air pressure.

Brazing (electrical resistance vacuum microbrazing) is an attractive means among many of rigid joining of the nozzle tip hydraulic cylinder group 4 to the insertion tip bore 2 capable of withstanding such cyclic distortions and collision impacts. Specifically, electrical resistance brazing in a vacuum, wherein a precise oxide-free thickness of probably a few microns of brazing material is deposited on one or both cylindrical surfaces to be joined such that the surfaces to be joined would fit together with a pressure contact when assembled by heating expansion of the outer cylinder, insertion tip bore 2, and freeze shrinking of the inner cylinder, nozzle tip hydraulic cylinder group 4, and sliding them together to a precisely positioned annular abutment shoulder of contact at or near the bottom of the nzzle body bore 2. When the temperatures equalize, the assembled parts will be in high pressure contact. Then heating electrodes are connected to the nozzle tip hydraulic cylinder group 4 and the nozzlebody bore 2, causing electrical heating at the interface seam to be brazed, melting the brazing material. In order to insure that after the brazing material is in a molten state, the nozzle tip hydraulic cylinder group 4 does not shift out of axial alignment with the insertion tip bore 2, centering positioning ridges of the original material of the nozzle tip hydraulic cylinder group 4 (typically steel) or the insertion tip bore 2 may be left as annular band or ridge shaped positioning spacers which come into contact with their opposite cylindrical steel wall either before or after the brazing material begins to melt. Thus the cylindrical brazed joint would exist in the gap(s) between at least two cylindrical bands of precision diameter concentrically positioning spacers (which contact their opposite cylindrical wall) carved of the originally higher surface(s) of one or both of the two cylinders to be brazed together.

For bypassing fuel around the nozzle tip hydraulic cylinder group 4, at or beneath the outer surface of nozzle tip hydraulic cylinder group 4, within the cylinder wall of proportionally about 400pm (0.4mm) in thickness (when FIG I is scaled to a 7mm nozzle insertion tip I diameter) are disposed numerous axially oriented relatively small cross section lower hydraulic chamber external channels 5A (being either surface grooves or holes drilled preferably by laser or spark erosion a short distance beneath its cylindrical surface), serving as conduits around the outside of the nozzle tip hydraulic cylinder group 4 for the fuel on its way to be com busted, which prior to its combustion functions as the hydraulic nozzle valve opening means. Making similar axial grooves or holes in or along the insertion tip bore 2 surface is less desirable due to the difficulty of such machining as well as weakening of the high pressure retaining insertion tip bore 2. As depicted in FIG I through FIG 4, which are accurately proportioned to represent a 7mm or a 9mm VCO nozzle valve (without limiting the present invention to any specific diameter), and for the purpose of the present example embodiment assumed to be specifically 7mm nozzle valves, these lower hydraulic chamber external channels 5A, although in size roughly on the order of the diameter of the nozzle holes 6 (typically about 100pm to 150pm in diameter (0.1mm -0.15mm) and 1mm in length), and three times their length, or 3mm, may amount to a collectively relatively large cross section fuel flow area disposed around the circumference of the 400pm wall thickness nozzle tip hydraulic cylinder group 4 in comparison with the cross section area of all of the nozzle holes 6, due to its relatively large diameter of about 3.3mm, having a circumference of it x 3.3mm = 10.4mm. Which 10.4mm circumference, if divided into 36 segments for 36 lower hydraulic chamber external channels 5A is three times as many as divide the face of a clock, each segment being 0.29mm, or 290 microns in width, of which the channel is 150 microns in width.

290pm = 150pm + 140pm, wherein the channel width is only lSOpm/290pm = 52% of the width of one of the 36 segments.

And if each lower hydraulic chamber external channel 5A were a hole of 200pm diameter (or equivalent area surface groove) separated by 9Opm within the 400pm thick cylinder wall, its flow improvement over the 150pm lower hydraulic chamber external channel 5A size is the ratio of their areas, or (200/150)2 = 178%. This arrangement is depicted in FIG 9, showing the cross section of the nozzle tip hydraulic cylinder group 4, being 4mm in diameter (4000 microns) with a cylinder wall thickness of 0.4mm (400 microns) having 36 holes of 0.2mm (200 microns) diameter each which can supply 12 times the cross sectional flow area of 12 nozzle holes of 0.1mm (100 microns) diameter (normal for contemporary fuel injectors). The purpose of this following arithmetic exercise, along with Drawing FIG 9, is to demonstrate that the 0.4mm hydraulic cylinder wall thickness is capable of conducting up to 12 times the flow-through required by the nozzle holes, and is therefore a realistic and robust design.

If there were to be 12 nozzle holes 6 of 150pm diameter (an unusually large number for such a large diameter, in practice), then the total cross sectional area of the 200pm diameter fuel channels surrounding the nozzle tip hydraulic cylinder group 4 is a cross sectional flow advantage of 36 x 1.78/12 = 5.33 times the cross section area of the 12 nozzle holes 6, although being 3 times longer in length, thus entailing more viscous flow resistance per channel.

For nozzle holes 6 of 100pm in diameter, which are currently popular due to the finer spray droplets they produce, the flow ratio advangage is eve more favorable, being 5.33 x (lSOpm/IOOpm)2 = 12 times the cross section area of all 12 nozzle holes 6 of 100pm diameter, 12 being an unusually large number of holes. Such advantageous nozzle tip hydraulic cylinder group 4 bypass flow ratios of 5.33 or 12 times the total nozzle flow areas (or flow rate capacities) having respectively 150pm or 100pm nozzle hole 6 diameters, would suggest adequate flow capacity (sufficiently low viscous flow resistance) to both quickly lift the valve closer unit 10 through a small valve lift of about 200 microns while at the same time injecting fuel into the combustion chamber without significant pressure drop due to the lower hydraulic chamber external channels 5A. If the 400pm wall thickness of nozzle tip hydraulic cylinder group 4 proves insufficient to withstand the pressure transients within the lower hydraulic chamber external channels 5A, the cylinder diameter may be reduced to provide a greater cylinder wall thickness.

This fuel emerges from the bottom ends of the lower hydraulic chamber external channels 5A, entering into the hydraulic valve lift control chamber 7 in order to execute both its fuel injection function and its hydraulic valve lifting function.

Said hydraulic valve lift control chamber 7 should, at its upper axial extremity (its axially upstream limit) provide a cylindrical shoulder (not shown) as a mechanical insertion stop to insure that the mechanical insertion of nozzle tip hydraulic cylinder group 4 into the insertion tip bore 2 axially positions said nozzle tip hydraulic cylinder group 4 correctly against said shoulder with the necessary high precision, which shoulder does not excessively block the lower hydraulic chamber external channels 5A. In order to avoid blocking the exit of the lower hydraulic chamber external channels 5A of the 200pm maximum diameter suggested for nozzle tip hydraulic cylinder group 4 of 4mm diameter, as well as to provide sufficient wall thickness for the 200pm diameter lower hydraulic chamber external channels 5A to withstand the fuel rail pressure within the 0.4mm (400pm) cylinder wall thickness of nozzle tip hydraulic cylinder group 4, said lower hydraulic chamber external channels 5A should be positioned exactly in the middle of said 400pm thick cylinder wall, leaving 100pm of material (eg steel) on both sides to contain the pressure, which is better than the corresponding ratio for the 7mm nozzle insertion tip 1 having a 4mm bore. In addition, because the pressure pulses are transient, wherein following the rising pressure transient the equilibrium pressure on both sides of said nozzle tip hydraulic cylinder group 4 may be the fuel rail pressure; and wherein following the falling pressure transient the equilibrium pressure on both sides of said nozzle tip hydraulic cylinder group 4 can be as low as atmospheric pressure, yet during the transient the pressure difference between the inside of said fuel channel 5 and and the surface outside the 100pm of material (eg steel) on both sides to contain the pressure may be at worst equal to the fuel rail pressure. Therefore, a mechanical insertion stop shoulder of an annular 100pm will avoid blocking said fuel channels 5 of 200pm diameter in the middle of said cylinder wall of nozzle tip hydraulic cylinder group 4 of 400pm thickness.

Drawing FIG 10 shows the present invention in the valve closed position equipped with Belleville springs instead of a conventional spiral spring as the valve spring means 3. The operational principle, and the essential parts of the embodiment of FIG 10 -12 and the previously described embodiments are functionally identical, retaining the same names and reference numbers, and operate identically as previously described.

The valve closer unit guidance radial support point 27 insures maximum precision of valve guidance and centering into the valve seat. Specifically, it is the inventive minimization of the distance of said point (or position) from the valve sealing contact surface of the valve seat which determines the stability and accuracy of operation of this embodiment, being in the present embodiment as drawn less than the diameter of the sealing ring of contact with the valve seat from the valve seat. It is also less than the diameter of the insertion tip bore 2, which likewise represents a novel degree of compactness. The close fit of the valve lift piston 9 in its cylinder comprises a secondary (lesser influence) constraint promoting valve guidance for the present embodiment.

The nozzle valve solution of FIG 10 may simply and directly substitute for the preceding nozzle valve solutions embodying spiral valve spring means 3 as incorporated in nozzle unit 28 (cf in FIG S & FIG 6) of the complete fuel injector of drawing FIGS, for example, resulting in considerable simplification below the level of the metering valve unit 29 of FIG 5, through substituting the axial fluid pipe 17 and its closely related spiral valve spring means 3 of FIG I -4, residing within the insertion tip bore 2, by its axially elongated (by comparison with FIG I -4) nozzle tip hydraulic cylinder group 4 (cf FIG 12), wherein the role of the annular transfer chamber 45 changes from feeding high pressure fuel into the nozzle valve to exhausting nozzle valve lift piston 9 pressure pulsations and leakage from the upper hydraulic chamber external channels SB back to the fuel tank via the low pressure supply channel 36 (cf FIG 12), exactly as it was described in the case of the coil spring based nozzle valve embodiment.

Referring to drawing FIG 12, in order to modify the embodiment of FIG S to accept the nozzle valve of FIG 10 in order to produce the embodiment of FIG 12, only the location of the metering valve outlet port 43 is changed in the fuel injector as depicted in FIG S from its top downward unto the boundary line between the metering valve unit 29 and the nozzle unit 28. The outline of nozzle unit 28 remains unchanged, but its switched flow supply channel 44 and low pressure supply channel 36 fluid channel connections appear to be reversed (cf FIG 12), but in fact preserve their previous functions.

Inflowing fuel from the switched flow supply channel 44 into the central trunk of the lower hydraulic chamber external channels SA is distributed via radially disposed branches of the same, each individually feeding the hydraulic valve lift control chamber 7. The low pressure supply channel 36 in FIG 12 is connected at the annular transfer chamber 4S to the upper hydraulic chamber external channel SB (shown in dashed lines also in FIG 10 symbolizing its presence in a plane containing the nozzle valve axis other than the cross section plane shown in FIG 10 -12), which enters the valve cylinder chamber 8 between a pair of radially disposed branching upper hydraulic chamber external channels SA, and there gathers its fuel either directly, or by aid of an optional flow equalization annular groove 58 in the top of the valve cylinder chamber 8 concentric with the nozzle insertion tip I axis so as to create a symmetric and uniform radial outflow and pressure distribution relative to the valve cylinder chamber 8 into the off-axis low pressure supply channel 36. If the optional flow equalization annular groove 58 is reduced in size or eliminated, hydraulic adhesion of the valve lift piston top 11 of the valve lift piston 9 to the cylinder top 12 will be increased, reducing the tendency for the valve closer unit 10 to bounce upon impact, as described near the end of the Disclosure of the Invention section.

Switched high pressure of fuel entering the hydraulic valve lift control chamber 7 hydraulically responds to a low pressure in the valve cylinder chamber 8, creating a pressure differential which lifts the valve lift piston 9 upwards until it contacts the cylinder top 12, which practically seals any further flow into the flow equalization annular groove 58, like in the case of FIG 1. This same high pressure similarly finds a channel of fluid communication to the nozzle holes 6 through the valve closer unit 10 bypassing the BeIleville spring stack valve spring means 3 as presented through the valve to piston shaft 59, namely the valve closer unit fluid channel 13, which routes a significant and possibly majority amount, if not all, of the fuel which finds its way to the nozzle holes 6 for injection into the combustion chamber when the valve lift piston 9 is lifted as stated, exposing the nozzle holes 6 to said fuel pressure. It will be noted that the valve closer unit fluid channel 13 passing through valve to piston shaft 59 has radially disposed inlet branches immediately below the bottom of the valve lift piston 9. Its outlet branches are similarly radially disposed as branches of the valve closer unit fluid channel 13, which have their outlet apertures in the surface of the valve tip section 20. The radial character of these channel branches is further indicated in the drawing by the two circles at the bases of the branches, at the valve axis, representing such branches perpendicular to the page. It is possible to dispose fuel channels also within the components of the valve spring means 3 for passing fuel axially through it.

By the same basic principle as was found operative in the embodiment of FIG 3, and applying its concluding paragraph to FIG 10 "The tiny volume of fuel ejected from above the valve lift piston top 11 when it rises to open the nozzle valve, which is the volume of valve cylinder chamber 8 (seen in FIG 1), is always sucked back into the same volume from whence it was expelled by action of the valve spring means upon the closing nozzle valve, which forces the nozzle valve closed and at the same time sucks this tiny amount of fuel back into its previous volume, the valve cylinder chamber 8. Therefore, in theory, the present system could operate at a zero net level of return fuel flow if there were no leakage losses past any of the slide bearings in the system, which is expected to be very nearly the case in practice".

For ease of assembly, and robustness of construction, the valve closer unit 10, which comprises the valve lift piston 9 and the valve tip section 20, and their joining valve to piston shaft 59, may consist of two parts as drawn, wherein the Belleville spring stack valve spring means 3 is assembled stacked upon the valve to piston shaft 59 by a close fit such as to insure precision of valve guidance and centering in its sealing seat.

In the present case (which comprises many possible variants not limited to those of earlier patents by the present author) which borrows the corrugated Belleville springs of an earlier patent application of the present author, PCT/EP2OIO/061790 found at 10,which does not include a functional example embodiment of an inwardopening opening nozzle valve, the only successful embodiments of that application having been outward opening nozzle valves, the parallel stacked Belleville springs have inner spacers 60 having an inner step, or shoulder (which is optional) such that the inner spacers 60 contact their neighboring inner spacers 60, and continuously (without gaps) contact the valve to piston shaft 59. The outer spacers 61 are of the same thickness as the inner spacers 60, are flat, and are flat spacer washers. The Belleville springs 62 are designed to have a contoured annularly corrugated relaxed shape of generally a single (or a double wave peak) wave peak as measured from the surface of the frustoconical section defined in space by their inner and outer edges by recursive finite element analysis to insure that their internal fatigue stress limits are not exceeded throughout their working stroke such that in their biased position, their inner and outer edges lie flat against their supporting spacer washers to allow their flat stacking at the inner and outer edges (or rims). This means that the Belleville springs 62 assume a curved shape when in their biased working position, wherein the curved cross section of the biased annularly curved Belleville springs 62 generally have two wave peaks.

In case the valve closer unit fluid channel 13 is insufficient to carry the injection flow and increasing its diameter becomes problematic (not currently foreseen), it is possible to perforate the Belleville springs 62 with holes or radial slots, or otherwise route additional fuel flow to the nozzle holes 6, for example through radial grooves in the Belleville springs 62, and/or their inner spacers 60 or outer spacers 61. Grooves in the wall of the insertion tip bore 2 are possible, but must reckon with weakening said bore against the injection pressure.

The valve tip congruent seating region 21, the valve seat congruent seating region 22, and the nozzle hole pressure equalizing groove 23 are as found in the embodiment of drawings FIG 3 and FIG 4.

Drawing FIG 11 shows the example embodiment of FIG 10 in the valve open position. As usual, in the interests of an unbroken narrative, the dynamic operation of the depicted nozzle valve from a closed to an open position was fully described in the previous section, for FIG 10. As in drawing FIG 1, in the nozzle valve open position the plane to plane hydraulically adhesive and valve bounce inhibiting contact of the valve lift piston top 11 with the cylinder top 12 collapses the valve cylinder chamber 8 volume to virtually zero. Notice that in the valve open position a claimed condition of the present invention's defining and enabling compactness is demonstrated, in that is contact point, also known as a valve opening limiting travel stop, is located at a distance of less than one insertion tip bore 2 diameter from a point of valve sealing contact at the valve seat. It is also within the same distance from a nozzle hole 6 outlet aperture, which are conditions which apply to all of the example embodiments of the present invention. Of course, it is possible to exaggerate, or stretch certain dimensional features of the present invention in order to circumvent these claims, and still gain a workable and substantial technical advantage of the present invention. Therefore such distortion is addressed by additional protective claims.

When the nozzle valve opens, the valve tip congruent seating region 21 is lifted axially to separate from its pressure sealing contact line(s) or surface(s) contained within the bounds of its grossly characterized "frustoconical" region from its mating contact line(s) or surface(s) contained within the bounds of the grossly characterized "frustoconical" valve seat congruent seating region 22, enabling fuel injection from the nozzle hole 6.

Drawing FIG 12 shows the embodiment of FIG 10 and 11 in combination with a complete piezoelectric fuel injector analogous to that of FIG 5 and FIG 6, demonstrating identical principles of operation common to both embodiments. The embodiment of FIG 12 purports to be an end of a complete fuel injector whose truncated upper part may be seen by reference to FIG 5.

Piezoelectric pushrod unit 41 initiates the fuel injection, opening the metering valve unit 29, admitting fuel from the high pressure supply channel 34, which accumulates in the metering valve spiral flow transfer annular cavity 26C and exits from the metering valve outlet port 43, and enters the switched flow supply channel 44, which enters the nozzle tip hydraulic cylinder group 4 of the nozzle unit 28 (which is close-fitted into the insertion tip bore 2 to minimize or eliminate fuel leakage), which fuel, by its pressure, lifts the valve closer unit 10, at the bottom of the nozzle insertion tip 1, causing a fuel injection.

As with the spiral spring embodiment of the present invention, a significant advantage of the present embodiment, is that a net flow of fuel is not spilled back to the fuel tank in order to operate the fuel injector, but rather a very limited amount of fuel leakage past the closely fitted sliding bearings of the moving parts of the present invention only is returned to the fuel tank via valve closer unit fluid channel 13 and low pressure supply channel 36, which flow is greatly minimized by the fact that the metering valve unit 29 is closed about 95% of an engine combustion cycle, during which time practically no leakage through said bearings should take place.

Threaded cap nut 39 retains the assembled internal parts of the fuel injector by threaded attachment to fuel injector body 38, the upper parts of which are not shown in FIG 12 being identical to the fuel injector of FIG 5.

An advantage of the embodiment of FIG 10 and 11 which becomes apparent in FIG 12 by comparison with FIG 5 is that due to the compactness of the Belleville valve spring means 3, the self-contained mechanism of the nozzle valve unit of FIG 12 is far more compact than even that of FIG 5, which enables the adaptation of the nozzle valve unit, including its valve spring means 3 into the tip of a standard 14mm spark plug, as claimed for the outward opening miniature nozzle valve application claimed by the present author in UK Patent Application 1104658.8, which increases the scope of applicability of that invention, as well as for any other application which may benefit from an ultra-compact inward-opening state of the art fuel injector.

Drawing FIG 13 may be viewed as a simple and modest modification of the embodiment of FIG 3 and FIG 4 wherein the nozzle driving high and low pressure supply connections have been reversed, along with two simple internal fluid channel changes. Because the embodiment of FIG 3 and FIG 4 is a minor modification of FIG I and FIG 2, the majority of description of the embodiments of FIG 1 to 4 apply obviously and directly to the present embodiment, and will not be repeated, although the same reference numbers are used in FIG 13 for reference to the corresponding descriptions for FIG Ito 4.

In FIG 13, the high pressure switched flow supply channel 44 (shown in FIG 14) is connected to and continues through the axial fluid pipe 17, and the low pressure supply channel 36 (shown in FIG 14) is connected to the insertion tip bore 2, in a simple reversal of the scheme for the embodiments of FIG I to 4, all else excepting internal nozzle valve hydraulic channeling remaining virtually identical to FIG I to 4.

Internal to the nozzle valve, high-pressure fuel impulses from the switched flow supply channel 44 from axial fluid pipe 17 enter into the valve closer unit fluid channel 13 of the valve closer unit 10 through the axial fluid channel coupling slide bearing 16 hydraulic and mechanical coupling unit of the spring base & piston channel unit 14B. The valve closer unit fluid channel 13 conducts these high-pressure fuel pulses through the valve closer unit to the hydraulic valve lift control chamber 7, where said high pressure acts upon the bottom of the valve lift piston 9 to lift it into the valve cylinder chamber 8 by unbalancing a lower pressure (provided by the low pressure supply channel 36 typically leading back to the fuel tank) in the valve cylinder chamber 8, which is vented to the insertion tip bore 2 through the orifice(s) of the upper hydraulic chamber external channel(s) 5B opening into the cylinder top 12 of the valve cylinder chamber 8, thereby collapsing the volume of the valve cylinder chamber 8.

The hydraulic principles of operation of this embodiment are identical to those of FIG I to 4.

Drawing FIG 13A shows a variant of drawing FIG 13 wherein instead of the high pressure pulses being conducted through the valve closer unit fluid channel 13 of valve closer unit to the hydraulic valve lift control chamber 7, the valve closer unit fluid channel 13 opens in the tip of the valve closer unit 10, below the nozzle holes 6 and therefore acts first upon the area of the tip of the valve closer unit 10 below its ring of valve sealing contact, wherever it may be located, as if the valve were an outwardly opening valve having the low pressure of the low pressure supply channel 36 on the other side of the ring of valve sealing contact, and not as if it were an inward-opening valve, wherein the high pressure which works to open he valve acts upon the circular area bounded by said ring of valve sealing contact to force it closed, rather than to lift it open. Therefore, the hydraulic valve lifting scheme of FIG 13A is more efficient than that of FIG 13. There may be a slight drawback, in that it is possible that this scheme of FIG I 3A may need a ring of valve sealing contact both above as well as below the ring of nozzle holes 6 in order to fully inhibit seepage of fuel to the nozzle holes 6. However, similar double rings of sealing contact are known and common in the current practice of the art.

Drawing FIG 13B has a valve closer unit fluid channel 13 which is a combination of that of FIG 13 and FIG 13A, having both kinds of orifices of valve closer unit fluid channel 13 present. The same hydraulic principles of operation and valve sealing as described for the embodiment of FIG l3Aapplyto FIG 13B.

Drawing FIG 14 is analogous to drawings FIG 6 and FIG 12, in depicting an enlarged lower section of drawing FIG 5 which shows the complete fuel injector from its electrical and fuel connecting end to its nozzle tip end. Likewise FIG 14 purports to be an end of a complete fuel injector whose truncated upper part may be visualized by reference to FIG 5.

ln FIG 14 may clearly be seen the nearly identical component parts and arrangement of the embodiment of FIG 6, but having the fuel channel connections to the nozzle unit 28 which appear to be identical to that of the embodiment of FIG 12-in otherwords, having its two fuel pressure connections, the switched flow supply channel 44 and the low pressure supply channel 36, reversed, along with subtle changes in the hydraulic channeling within the nozzle valve to agree with the reversal of the pressure connections, as discussed for FIG 13.

The more comprehensive view of FIG 14 may help to appreciate certain possible advantages or differences with the arrangement of FIG 6. One such difference which may in some cases be an advantage is that the high pressure channel 44 is comparatively extremely restricted and severely contained in the embodiment of FIG 14. In view of the challenges of higher operating fuel pressures in the progress of the art, and the demands of providing more robust pressure containment nozzle structures for this application, this difference may be decisive in some designs.

Drawing FIG 15 shows an embodiment of the present invention which eliminates the need for a return flow connection to the fuel tank or other storage vessel for fuel seepage from the fuel injectors. It enables a single hydraulic connection to the fuel injector, instead of two hydraulic connections.

Recall that the amount of fuel returned to the fuel supply (eg the fuel tank) in the present invention consists only of fuel seepage past sliding bearings in the fuel injectors, and does not include deliberate spillage of fuel. Such deliberate spillage of fuel is common in the state of the art for operational reasons, and creates substantial technical challenges, such as waste of heat energy via the returned fuel, cooling of that fuel, evaporation of that fuel, etc., and the only technical solution to eliminate this problem is a non-return fuel injector design requiring direct actuation of the nozzle valve by the piezoelectric unit, which has technical challenges not present in the present invention, offering a very distinct technical alternative solution, with comparable if not superior benefits and advantages.

The embodiment of FIG 15 is essentially the same as similar to the embodiment of FIG 13 and FIG 14 and most specifically like the embodiment of FIG ISA 13B which provides high pressure fuel impulses from the switched flow supply channel 44 at the outlet orifice of valve closer unit fluid channel 13 at the tip of valve closer unit 10 and branches of valve closer unit fluid channel 13 discharging into the hydraulic valve lift control chamber 7 with the exception that the return flow path of the embodiment of FIG l3Ato the low pressure supply channel 36 (cf FIG 14) via the upper hydraulic chamber external channel SB has been eliminated. The difference is in that a low pressure creating valve closer unit venturi 63 is provided in the present embodiment to replace the function of this return flow path to the low pressure supply channel 36, which provides a source of low pressure within the valve cylinder chamber 8 to unbalance the valve lifting high pressure supplied to the hydraulic valve lift control chamber 7 and to cyclically expel from and suck back into the valve cylinder chamber 8 its fuel volume plus a minute amount of seepage through the sliding bearings to the low pressure supply channel 36.

In the present embodiment, the valve closer unit venturi 63 is disposed within and as part of the axially disposed valve closer unit fluid channel 13 near the tip of the valve closer unit 10, wherein are provided at least one, but preferably a multiplicity of radially symmetrically disposed and beneficially angled fluid channeling branching connections of the valve closer unit fluid channel 13 between the valve cylinder chamber 8 and the valve closer unit venturi 63. Said beneficial angle of fluid channeling connection is such that for the stream of fluid passing normally through the valve closer unit venturi 63 to flow into the connecting branch of valve closer unit fluid channel 13 connecting with the valve cylinder chamber 8, would require a partial reversal of direction of the normal stream of fluid flow, and a reversal of its momentum. This is also the most beneficial angle of connection of branch channels of a venturi, to accomplish its primary function of creating a lower pressure in said branch channels than at the exit orifice of the venturi, which in the present case will be the volume of the hydraulic valve lift control chamber 7 which comes into fluid communication with the exit of the valve closer unit venturi 63 at the moment when the fluid flowing out of the nozzle holes 6 creates enough of a pressure drop in the branch channels of the valve closer unit venturi 63 connected to the valve cylinder chamber 8, in contrast to the pressure in the hydraulic valve lift control chamber 7, to expel valve lift blocking fluid from the volume of the valve cylinder chamber 8 into the valve closer unit venturi 63 to begin to lift the valve closer unit 10 against its valve spring means 3. Once this valve lifting process begins, a positive feedback effect in conjunction with progressively increasing fuel flow through the valve closer unit venturi 63 accentuates and reinforces this lifting process.

It is important to recognize that the venturi effect is not dependent upon a the illustrated decreased (constricted) diameter of the venturi at the connecting point of its branch channel(s) in order to produce a reduced pressure in the branch channel(s) with respect to that at a point of stationary fluid in fluid communication (eg valve cylinder chamber 8) with the moving fluid at said branch connecting point. Thus, the venturi can be of constant diameter, without a narrowing section, for it to work properly and effectively due to tt1i phenomenon Bernoulli's principle of decreased pressure within a volume of a moving stream of fluid in comparison with a stationary, or lower velocity region of fluid, wherein there is fluid communication between said parts of the fluid at differing velocities. Because such narrowing increases the venturi effect pressure drop, and is characteristic of many venturis, the valve closer unit venturi 63 has been illustrated with such a narrowing section, and not because it is required for the present invention.

It is intuitively evident that if injection pressure is supplied to the nozzle valve in its open position, the negative pressure of maximum flow through the venturi will be capable of maintaining this valve in the open position, provided that the valve spring force is not excessive, or provided an enormous injection pressure capable of overwhelming any reasonable nozzle valve spring, which at today's 2500 bar fuel rail pressures moving steadily beyond 3000 bar is the present reality.

The forces acting to close the valve closer unit 10 are the force of the valve spring means 3, the injection pressure acting upon the area of the bore of the piston coupling and guide pipe bearing 19, and the pressure within the valve cylinder chamber 8, which originates in the valve closer unit venturi 63. This last pressure may be above or below atmospheric depending on the rate of flow through the venturi.

If radial branches of the valve closer unit fluid channel 13 did not connect to the hydraulic valve lift control chamber 7 (as is the case for the embodiment of FIG 15B). the forces acting to open the valve closer unit 10 would initially be the injection pressure acting upon the circular contacting area between the conical valve seat and the conical valve tip, which for the present invention make contact above the nozzle holes 6, but below the practically coincident upper edges of the valve tip congruent seating region 21 and the valve seat congruent seating region 22. It is clear that this circle of contact is larger than the bore hole of the piston coupling and guide pipe bearing 19 causing a downward force due to the injection pressure on its circular area. The problem is that in the absence of sufficient flow velocity through the valve closer unit venturi 63 before the nozzle valve opens, sufficient to lift the valve closer unit 10 against its valve spring means 3, the pressure in the valve cylinder chamber 8 will instantly attain the injection pressure from the switched flow supply channel 44 (see FIG 16), which will force the nozzle valve shut.

At issue is the effectiveness of the venturi from the very start of the injection to reduce the pressure in the valve cylinder chamber 8. To solve this issue, one must examine drawing FIG 20, 20A, and 20B. This drawing shows the tip of a "normal" VCO nozzle similar to that of Drawing FIG 22 and Drawing FIG 1, whose geometry below the level of the nozzle holes is identical for all embodiments of the present invention. Drawing FIG 20 does not show an angular difference between the conical needle tip and its approximately conformal valve seat, which is normal, and assumed to be present for the present invention. Normally, sealing contact occurs at a circular ring of contact above the nozzle holes due to this small angular difference. With wear, the ring broadens into an annular band. Various strategies are known embodying multiple rings of sealing contact.

FIG 20B shows that in addition to the described (not drawn) small angular gap between the needle tip and its conical seat, a deliberate gap is created between the upper edge of the nozzle hole pressure equalizing groove 23 and the lower edge of the inlet orifice of the nozzle holes 6 after hydro-erosive grinding of its edge to insure the fastest fuel injector response to an electronic injection signal without suffering leakage of fuel from the nozzle holes -relying on capillary action of the fuel to prevent this. This gap should not contribute to fuel evaporation between combustion events, because the fuel in the nozzle hole pressure equalizing groove 23 is retained by capillary action, and protected from combustion heat by the cyclically fuel cooled thermal inertia of the valve closure element 10. The combination of these two gaps is substantial, and it is precisely in the closed position of the nozzle valve that the pressure boosting effect at the inlet apertures of the nozzle holes of the pressure equalizing groove 23 may be critical for providing sufficient immediate flow through the valve closer unit venturi 63 at the instant before the nozzle valve unseals. The smaller the needle to seat gap, the larger is the relative pressure boosting benefit of the pressure equalizing groove 23.

Given the pressure equalizing groove 23 in combination with a sufficient needle to seat gap, and a sufficient gap between the upper edge of the pressure equalizing groove 23 and the lower edges of the nozzle holes 6 for the purpose of providing instant flow through the valve closer unit venturi 63, that venturi may be capable of generating negative pressure immediately in the valve cylinder chamber 8, which, according to the above analysis may produce sufficient upward force to overcome any practical valve spring force, due to use of overwhelmingly large modern standard fuel injection pressures, now routinely at 2500 bar and going higher.

In addition to the valve closer unit venturi 63, FIG 15 shows at the lower tip end of the valve closer unit 10 two additional elements which should assist the venturi in providing maximum lifting effect and injector efficienct Maximum lifting effect before the nozzle valve begins to open requires full injection pressure to be present in the hydraulic valve lift control chamber 7, which is supplied by the radial branch channels of the valve closer unit fluid channel 13 immediately at the exit of the valve closer unit venturi 63. This feature is labeled in the enlargement, FIG ISA. This injection pressure in the valve lift control chamber 7, in combination with the same pressure coming from the tip of the valve closer unit 10, together exactly balance the initial injection pressure in the valve cylinder chamber 8 in combination with the injection pressure at the upper aperture of the piston coupling and guide pipe bearing 19. This leaves two remaining forces acting upon the valve closer unit 10, the valve spring means 3 force and the venturi force acting within the valve cylinder chamber 8. Given that the venturi force is provided by the immense nearly 3000 bar fuel injection pressure one should expect a reasonable possibility that the venturi will win.

We note that the radial branch channels of the valve closer unit fluid channel 13 immediately below the valve closer unit venturi 63, which feed the hydraulic valve lift control chamber 7, emerge from an enlarged diameter cylindrical cavity, whose diameter is substantially larger than that of the venturi. The purpose of this cavity is to substantially reduce the Bernoulli effect (the venturi effect) in these radial branch channels.

Finally, we observe the obvious check valve at the outlet orifice of the valve closer unit fluid channel 13 at the bottom tip end of the valve closer unit 10, whose purpose is to insure that the large fuel volume between this check valve and the mushroom valve 40 of the metering valve unit 29 (see FIG 16) would not leak out of the nozzle holes 6 due to the normal angular gap between the valve closer unit 10 conical tip and its valve seat, and the gap opening into the nozzle holes 6 due to hydro-erosive grinding illustrated in FIG 20B.

Although the vacuum of the closed volumes involved, in combination with the capillary effect should make such a check valve unnecessary, it is provided in case of need.

Practitioners of the art will choose the type of check valve most suitable for their application, in lieu of that which is illustrated. Ceramic balls may be preferable, being much lighter in weight and much more durable than their steel counterparts.

FIG 15 presents the most probably successful of the venturi-based or any of the purely non return flow embodiments (no low pressure supply channel 36) of the present invention.

Drawing FIG 15A provides an enlarged view of the tip of the valve closer unit 10, showing the valve closer unit venturi 63, the radial branches of the valve closer unit fluid channel 13 as they emerge from an enlarged diameter cylindrical section of the axial trunk of the valve closer unit fluid channel 13, whose enlarged diameter is provided to reduce the Bernoulli effect upon these radial branches due to the flow through the axial trunk of the valve closer unit fluid channel 13. And sealing the exit orifice of the valve closer unit fluid channel 13 is the check valve, designed to prevent fuel leakage from the nozzle holes 6 between injections.

Drawing FIG 15B illustrates an increase of the nozzle valve sealing contact ring diameter for the purpose of increasing an initial hydraulic valve lifting advantage in the absence of radial branches of the valve closer unit fluid channel 13 serving the hydraulic valve lift control chamber 7. The original lifting force is provided by the pressure from the nozzle tip acting upon the area of this circle, in opposition to the valve spring pressure, the injection pressure acting upon the bore of the piston coupling and guide pipe bearing, and the pressure with in the valve cylinder chamber 8 acted upon by the venturi as described in the preceding section. Once the nozzle valve breaks its closing seal, After the sealing contact of valve closure is broken, the hydraulic valve lift control chamber 7 almost immediately achieves full injection pressure, greatly increasing the opening force of the nozzle valve.

The principle of this drawing is applicable to several embodiments of the present invention Drawing FIG 16 illustrates how the venturi-based embodiments of FIG 15, ISA, and 15B may be integrated into a complete fuel injector, being is analogous to drawings FIG 6, FIG 12, and FIG 14, in depicting an enlarged lower section of drawing FIG 5 which shows the complete fuel injector from its electrical and fuel connecting end to its nozzle tip end.

Likewise FIG 16 purports to be an end of a complete fuel injector whose truncated upper part may be visualized by reference to FIG 5.

In FIG 16 may clearly be seen are found the nearly identical component parts and arrangement of the embodiment of FIG 14, and operation by practically the same principles, with the difference being the replacement of the low pressure supply channel 36 as the source of low pressure provided in the valve cylinder chamber 8, which is necessary for the lifting of the valve closer unit 10, by the valve closer unit venturi 63. Therefore the low pressure supply channel 36 has been eliminated, which also permitted (but did not demand as essential) the elimination of the slender part of the axial fluid pipe 17 which passed through the coils of the valve spring means 3, as well as the elimination of its coupling structure with the valve closer unit 10, the axial fluid channel coupling slide bearing 16, from the spring base & piston channel unit 14B of FIG 13, resulting in a new reduced structure, the spring base & piston channel unit 140, and the truncated axial fluid pipe 17A.

Please bear in mind that for the present embodiment it is optionally equaHy valid to retain the original axial fluid pipe 17 and spring base & piston channel unit 14B exactly as shown in FIG 13. What is essential for the present embodiment is the elimination of the low pressure supply channel 36 and its connection to the nozzle valve via the annular transfer chamber 45. Indeed there may be circumstances in which the option of retention of the full axial fluid pipe 17 along with its mating spring base & piston channel unit 14B as shown in FIG 13 is preferable, such as employment of exceptionally high fuel pressures, or where the compressibility of the relatively larger fuel volume within the insertion tip bore 2, or vibrational or wave phenomena affecting the injection process become problematic.

Of course, elimination of the low pressure supply channel 36 also eliminates its associated fuel return channel connection 31, which is for clarity of function illustrated as a fuel hose nipple in FIG 5. Broadly applied, FIG 5 substantially represents all of the embodiments of the present invention including that of FIG 16, when the relatively minor or progressive or optional variational differences each embodiment entails are taken into consideration.

Drawing FIG 17 applies the venturi principle thus far applied in the transformation of the embodiment of FIG 13 to the embodiment of FIG 15 to similarly transform the embodiment of FIG 10 to the embodiment of FIG 17, eliminating the low pressure supply channel 36 and replacing its function of enabling the low pressure collapsing of the valve cylinder chamber 8 by application of a valve closer unit venturi 63. Indeed this principle of replacement of a low pressure or atmospheric pressure source, such as a fuel tank by a venturi may be broadly extended to applications which are impractical to individually detail in drawings, such as for example one or more independent and external venturis attached to the high pressure fuel pump, and serving the low pressure sourcing needs of multiple fuel injectors in an engine, but may be abstractly described, and are thus claimed in the present invention.

The venturi arrangement of the embodiment of FIG 17 is in principle the same as that of FIG 15 wherein fuel approaches the nozzle holes 6 from below, namely in the direction of increasing valve seat radius (as in the case of an outward opening valve, although the present valve is in fact an inward opening valve), from the outlet orifice of the valve closer unit fluid channel 13 through the axis and at the tip of the valve closer unit 10: the upper branches of the valve closer unit fluid channels 13 which function as low pressure generating branch channels of the valve closer unit venturi 63, having orifices opening into the valve lift piston top 11 of the valve lift piston 9 of the valve closer unit 10 generate at their angled connection with the flow-through bore of the valve closer unit venturi 63 the relatively lower pressure supplied to the valve cylinder chamber 8 at the top of the valve lift piston 9 by the valve closer unit venturi 63 branch channels, to unbalance the respectively higher pressure of the hydraulic valve lift control chamber 7 lifting by its higher pressure the bottom of the valve lift piston 9, which pressure is supplied by the relatively higher output pressure of the main flow-through channel of the valve closer unit venturi 63. As the valve closer unit venturi 63 generated high pressure from below the valve lift piston 9 (in hydraulic valve lift control chamber 7) responds to the valve closer unit venturi 63 generated low pressure in valve cylinder chamber 8, the resulting pressure differential force works against the valve spring means 3 to lift the valve lift piston 9 into the volume of the valve cylinder chamber 8. And this collapsing volume of relatively incompressible fuel is thereby ejected into the main stream flowing through the valve closer unit venturi 63 at a lower pressure than that forcing the valve lift piston 9 upwards, and which is ejecting said fuel volume, most of which flows out through the nozzle holes 6.

When the injection ceases, and the valve spring means 3 retracts the valve lift piston 9 from the valve cylinder chamber 8, creating a void, fuel is sucked into said void from the valve closer unit fluid channel 13 angled branch connection with the valve closer unit venturi 63 until the moment when the valve tip congruent seating region 21 makes seating contact with the valve seat congruent seating region 22.

The low pressure creation functionally advantageous direction of said angled union of the valve closer unit fluid channel 13 branch channels of the valve closer unit venturi 63, being the same as in the embodiment of FIG 15, is such that as the main, axial venturi fuel flow approaches said union of the main channel of the valve closer unit venturi 63 with its angled branch channels, said branch channels radially approach the axis of the valve closer unit venturi 63, which is the axis of the valve closer unit 10.

Said valve closer unit venturi 63, including its low pressure producing venturi branch channels which are a part of the overall valve closer unit fluid channel 13, of which in fact the valve closer unit venturi 63 is a part, is disposed advantageously mostly within the bounds of the valve lift piston 9, which is mostly axially above the level of the Belleville spring stack valve spring means 3.

A dynamic mechanical and hydraulic connection of the valve closer unit 10 with its supply of pulsed high pressure fuel of the switched flow supply channel 44 is provided by the axial fluid channel coupling slide bearing 16 making sliding bearing contact inside the bore of the piston coupling and guide pipe bearing 19 connected to the switched flow supply channel 44, which axial fluid channel coupling slide bearing 16, as in the previous embodiments, functions both to guide the valve closer unit 10 centered into its valve seat in alignment with the valve axis, while also providing (in the present case by its outer sliding bearing surface) a leakage resistant sliding bearing hydraulic coupling channeling the high-pressure fuel from the switched flow supply channel 44 to the nozzle holes 6.

An optional alternative location for the valve closer unit venturi 63 other than described above is provided within a venturi cylinder 64 is shown by dotted lines disposed at the top of the valve cylinder chamber 8 provided low pressure generating fluid connections via the upper hydraulic chamber external channels SB opening into the cylinder top 12 to the valve closer unit venturi 63 feeding its main discharge into the piston coupling and guide pipe bearing 19, such that the lower surface of the venturi cylinder 64 is the upper surface of the valve cylinder chamber 8, being namely the cylinder top 12. It is possible to operate both venturis at the same time.

This particular implementation of the venturi, namely of the venturi cylinder 64, particularly illustrates the broad possibilities of disposing such a venturi practically anywhere within the hydraulic path of high pressure departing from the high pressure fuel pump of the fuel injection system. Notwithstanding the advantages of the venturi locations of the illustrated embodiments of the present document, venturi locations nearer to the fuel pump at any point along its high pressure fuel path may be chosen for any purpose, including the presently described purpose, required to create a source of lower pressure than within said high pressure fuel path into which leakage flow, or spillage flow, for example, created by said high pressure fuel path, may be returned through said angled branch connections of said venturi to the main stream of said high pressure fuel path, without having to return this leaked or spilled fuel back to the fuel tank. A venturi disposed in such a manner in effect, "mops up", or absorbs such high pressure spillage, or the cyclical back and forth flow of a fixed small fuel volume, as that of the valve cylinder chamber 8, avoiding the need for such spillage or cyclical flow to be disposed or vented outside of the high pressure system, for example by return to the fuel tank.

One of the more obvious such alternative locations for the venturi cylinder 64 could easily be in the large mostly unoccupied lower 1/3 part of the cylindrical metering valve unit 29 in FIG 16 (disposed above the nozzle unit 28 and its nozzle insertion tip upper part IA hosting the valve cylinder chamber 8 to be served by said venturi), for example, whose sole purpose has been, as illustrated, to channel the switched flow supply channel 44, which switched flow section of the supply channel 44 could be an ideal alternative location within the switched flow supply channel 44 for hosting the valve closer unit venturi 63 as part of the venturi cylinder 64, whose sections of the upper hydraulic chamber external channels 5B could be extended in length to reach and open into the valve cylinder chamber 8 within the nozzle insertion tip upper part IA.

Finally, a full drawing of the embodiment of FIG 17 as a complete fuel injector, similar to FIG 5, or even FIG 16 is not necessary, because the operating principles of the embodiments of FIG 15 and FIG 17 are identical, and therefore a full depiction of the complete fuel injector embodiment of FIG 17 may be visualized by reference to FIG 16, by removing the coil valve spring and its truncated axial fluid pipe 17A threaded spring retainer, and replacing these, along with the unnecessarily large diameter insertion tip bore 2 by the much smaller diameter axial channel, the switched flow supply channel 44 as indicated in FIG 17, as extending to the top of the nozzle unit 28, whose only internal feature above the piston coupling and guide pipe bearing 19 is, in the case of the larger view of the embodiment of FIG 17, the switched flow supply channel 44. All else is the same as shown in FIG 16, where in the counterpart of FIG 5 the fuel return channel connection 31 (the hose nipple) is eliminated.

Of course the benefits of the valve closure element 10 tip modifications shown in FIG 15 ISA, or 15B are valid for the embodiment of FIG 17 Drawing FIG 18 (along with the related embodiments of FIG 19 and 19A) shows that in principle, the preceding ventur-based embodiments stripped of their venturis and slightly modified shoUld provide at least elementary functionality as fuel injectors, but may be less effective in providing adequate valve lift in all circumstances, subject to future study. The embodiment of FIG 18 is nearly identical to that of FIG 15B, or the lower end of FIG 16.

The difference is the present absence of the venturi, plus the unification of the hydraulic valve lift control chamber 7 with the valve cylinder chamber 8 by means of interconnecting pressure equalization channels 65. This eliminates the valve lifting effect of the valve lift piston 9, whose function now is reduced to valve guidance by its bearing surface, and the valve travel stop function of the opposite inner surfaces of the valve cylinder chamber 8.

The hydraulic equivalent of the embodiment of FIG 18 is shown in FIG 19 and 19A, where non-essential elements have been stripped away, reducing the mechanism to its essentials.

The mechanism of FIG 18 should function based on the differential force resulting upon application of hydraulic injection pressure in the insertion tip bore 2, as follows: Applied injection fluid pressure acts upon the circular area of the piston coupling and guide pipe bearing 19 to produce a downward force, closing the nozzle valve. This same hydraulic pressure also acts upon the circle of valve sealing contact disposed above the nozzle holes 6, but below the upper edges of the valve tip congruent seating region 21 and the valve seat congruent seating region 22. The area of this circle is clearly larger than the area of the former circle (the cross section of the piston coupling and guide pipe bearing 19), resulting in a net upward force acting against the valve spring force, which is downward. This valve spring force can be exceeded given sufficient injection pressure, which for today's 2500 bar to 3000 bar pressures should not be challenging.

The only issue with this design is that higher injection pressures may be needed to produce equivalent injection performance (fuel atomization).

Drawing FIG 19 illustrates the embodiment of FIG 18 stripped of its non-essential features. leaving only what is essential for equivalent functionality, reducing the inertial mass of the system.

Drawing FIG 19A illustrates the embodiment of FIG 19, wherein the diameter of valve sealing contact has been increased, and consequently the valve lifting differential force of the embodiment of FIG 19 has been thereby improved. The drawback of this approach is reduced valve sealing pressure, requiring a stronger valve spring, and therefore greater lifting force to open the valve.

Drawing FIG 20 illustrates a non-essential, but practically indispensable feature of all of the embodiments of the present invention, the nozzle hole pressure equalizing groove 23, whose effect is to immediately increase the fuel pressure at the inlet orifices of the nozzle holes 6, resulting in better fuel atomization, and greater volume rate of fuel delivery. This feature is an ideal complement to the inherent valve travel limiting ability of the present invention, enabling the valve to be stopped without bouncing (without rebound) at precisely its optimum vaive opening position or level. This is a dramatic improvement of the state of the art, with the possible exception of directly actuated piezoelectric driven nozzle valves, which have their own unique problems. The nozzle hole pressure equalizing groove 23, according to PCT/1B201 2/051770 also by the present inventor, should substantially improve the fuel injection performance of all of the embodiments of the present invention.

FIG 20 illustrates that the effect of increased fuel pressure at the nozzle holes begins at the instant of initial opening of the nozzle valve, causing a sharp rise to full fuel pressure at the nozzle valves much sooner than without benefit of said groove, resulting in the possibility of using much shorter injection pulses, at much higher injection repetition rates per combustion cycle.

Drawing FIG 20A illustrates what is anticipated to be an optimal positioning of the upper edge of the nozzle hole pressure equalizing groove 23 with respect to the lower edge of the nozzle holes 6, resulting in both optimum protection of the trapped fuel volume within the nozzle hole pressure equalizing groove 23, as well as instantaneous pressure response at the nozzle holes 6 upon energization of the fuel injector (no injection delay, with nearly instantaneous pressure rise to maximum). The same fast response applies also to de-energization.

Drawing FIG 20B illustrates the result of hydro-erosive grinding of the inlet orifice edges of the nozzle holes 6, resulting in a small visible gap between the upper edge of the nozzle hole pressure equalizing groove 23 and the lower edges of the nozzle holes 6. Capillary action along with the vacuum which retains trapped fluids in their trapped volumes despite small openings to the atmosphere is expected to prevent leakage from this gap. as well as from the normal angular gap present between the conical valve tip and valve seat in virtually all inward-opening fuel injectors, whereby a sealing ring of contact is established for valve sealing by use of two slightly different cone angles for the two sealing surfaces.

This gap is present in all VCO nozzles, noted for their superior resistance to fuel leakage (in contrast to SAC nozzles), demonstrating the principle which should ensure similar superior performance in the case of FIG 20B.

Drawing FIG 21 illustrates by a single example the multiple fluid (liquid or gas) injection capability of all of the embodiments of the present invention, wherein any number of metered fluid source channels (piped metered fluid sources) are united to feed into a single receiving channel, the combined switched flow supply channel 70, by means of check valves (one-way valves), all allowing flow from said fluid sources only into the single combined switched flow supply channel 70, which feeds injection pulses into the nozzle valve.

The present invention is adaptable to multiple fuel operation, capable of virtually simultaneous injections of multiple liquids and gases from its single nozzle valve due to its exceptional high speed, high frequency injection capability, wherein multiple metering valves or pumps may be connected to the same fuel injector nozzle valve inlet channel by use check valves.

The exceptional high speed, high frequency operation of the present invention enables precise and extremely finely, precisely metered digital control of injection of multiple fluids such as gaseous fuels, liquid fuels, and water possibly tens of times into one and the same combustion event (cycle), thus exerting unprecedented control over the internal combustion process in motor vehicle operation, for performance, economy, and reduced undesirable engine emissions, and quieter and smoother engine operation.

Thus the two fluid sources shown in FIG 21 are the switched flow supply channel 44, and its analogous switched (metered) alternate gas or liquid supply channel 67. The high pressure supply channel 34 with its associated high pressure connector, the high pressure input pipe coupler 30 as shown in FIG 5 is metered by the metering valve unit 29 to produce as output feeding the nozzle valve, the switched flow supply channel 44. And the switched (metered) alternate gas or liquid supply channel 67 is produced similarly, either internally within the body of the fuel injector, or externally, for which purpose many commercial metering valves for diverse gases and liquids are readily available for connection to the present invention.

The embodiment of FIG 21 well illustrates the application of compressed natural gas as the principal fuel for powering diesel engines. It is well known that compressed natural gas can not be ignited by compression ignition, but it can be ignited by a preceding short injection of diesel fuel, whose combustion in the diesel engine combustion chamber is immediately utilized as the initiator of combustion for a subsequent compressed natural gas injection, providing the vast majority of the combustion energy. Thus, the diesel portion of the present system would consist of the metering valve unit 29 with its switched flow supply channel 44 metered diesel fuel output being fed into the primary fuel check valve 68. A second metering valve, not shown, meters compressed natural gas into the switched (metered) alternate gas or liquid supply channel 67, which feeds into the alternate gas or liquid check valve 69. The metered diesel fuel at its check valve is approximately three times the pressure of the metered gaseous fuel at its check valve. This is very convenient for the sake of lubrication of the present invention in cases where at least one of any combination of alternate fluids is non-lubricating (water, gas, liquified gas, etc.), because application of at least one diesel fuel injection through the system per combustion cycle should suffice to maintain adequate Jubrication of the system. The pilot ignition diesel injection serves this purpose. Furthermore, since it is desirable to start the time critical pilot diesel injection precisely, the preceding injection of gas must have been previously purged from the system. Due to the high speed, high frequency operation of the diesel injection part of the system, an ultra short microinjection of diesel fuel can be applied to purge the gas at the terminus of the gas injection without injecting a substantial amount of diesel fuel into the cylinder, in preparation of the following short diesel pilot injection.

Water injection has proven beneficial for engine operations, and the present capability will open up previously unexplored possibilities for improvement of engine operation.

Drawing FIG 21A shows an enlargement of the ball check valve. Such balls may advantageously be ceramic due to their light weight and low wear properties.

Drawing FIG 18 22 shows a classic VCO nozzle insertion tip I typical of the 7mm and 9mm standards for convenience of comparison with the advantages of the present invention. It is shown in the valve closer unit 10 (being for the art, and not for the present invention a "needle valve") open position, which for the scale of a 7mm VCO nozzle valve would correspond to a 60 microns valve lift, being a small valve lift according to the state of the art where 200 to 350 microns is normal (and far more than is required to establish maximum pressure and flow at the nozzle valves, due to the impossibility of using valve closer unit 10 travel stops due to valve bounce in the absence of the benefits of the present invention (nevertheless shown here for illustration purposes only), but compensated for that fact by being furnished with an nozzle hole pressure equalizing groove 23 lifted to the same axial level as the outlet orifices of the nozzle holes 6, which raises the fuel pressure at said orifices to compensate for the low valve lift. Also, the valve tip congruent seating region 21 is shown in correct raised relationship to the valve seat congruent seating region 22. Again the reminder that no specific valve seat containing annular area or areas can be specified within said congruent seating regions in any of the drawings of the present patent document because current art places, in its patents it not in its commercial practice, one or more such circles or annular bands of valve sealing contact in various places, in various numbers, and for various purposes, and with various surface contours.

Description of the invention in terms of its specific claims Below are quoted the specific claims of the present invention for the purpose of description of the present invention. In general, these claims quoted immediately below from the Claims section, may be regarded as a concise and compact summary of those unique attributes presented in the preceding Description and the Drawings which are distinct from, or not found in, the state of the art. Comments are added to these claims as needed, clarifying their relationship to the preceding part of the Description and/or the Drawings.

Claim 1. A direct injection inwardly opening differential valve type nozzle" based fuel injector, otherwise characterized as operating in a broad sense (allowing for inventive improvement) by the principle of,,nozzle differential ratio", often generically termed hydraulic" (also known as a,,mechanical"), depicted for purposes of reference orientation with its nozzle insertion tip I tip pointing for purposes of descriptive reference and orientation in the relatively "downward" direction, comprising a nozzle insertion tip I (optionally divisible into sections such as nozzle insertion tip upper part IA and nozzle insertion tip lower part I B) extending axially toward the combustion chamber beyond a point of combustion chamber pressure sealing surface contact of said injector with the engine (such as for example an annular shoulder of increased diameter, often sealed by a copper gasket, or tapered pipe threads, etc.) or an intermediary gasket, for direct injection communication penetration of a combustion chamber pressure containment structure (for example, a cylinder head thickness, sometimes comprising cooling elements such as liquid coolant passages or air cooled fins, etc.); wherein a nozzle valve closer unit 10 comprises an hydraulic valve lift piston 9, and a valve tip 20 which in turn comprises a generally tapering valve tip congruent seating region 21, where in turn is comprised one or more valve sealing contact surfaces establishing valve sealing contact with a valve seat congruent seating region 22; which nozzle valve closer unit 10 is operated (displaced) by hydraulic pressure comprised within a valve lift control chamber 7 controlled by a pulsed fuel flow controlling (metering or dispensing) means, (for example an injection pump, or a valve, connected to a fuel pressure source such as a fuel rail, or a pump, or their combination, etc.) and which may be integral and/or external to the hydraulic fuel injector; characterized by the sliding piston bearing surface (establishing said piston operative pressure differential) of the valve lift piston 9 of said valve closer unit 10 being limited by being confined in its entirety within the nozzle insertion tip (1); and its valve spring means 3 being limited by being confined, to at least part of its axial length, within the nozzle insertion tip 1.

Comment: This is the most general and broadest attribute of this aspect of the present invention (pertaining to the present one of a plurality of independent claims within an unity of the present invention), which is evident in all of the drawings and embodiments of the present invention showing the nozzle insertion tip 1.

Claim 2. The fuel injector according to claim 1, wherein the entire inertial (moving) mass of valve closer unit 10 is confined to the downstream (tip end) half of the nozzle insertion tip I length.

Comment: This attribute is evident in every drawing of example embodiments of the present invention showing the nozzle insertion tip I in combination with the dimensions of said tip 1 stated in the preceding part of the Description, but is fully graphically evident in FIG 5,6, 12, 14, and 16.

Claim 3. The fuel injector according to claim I or 2, wherein the sliding piston bearing surface (establishing said piston operative pressure differential) of the valve lift piston 9 of said valve closer unit 10 being limited by being confined within the downstream (tip end) one third (113) of the nozzle insertion tip I axial length, or entirely within one nozzle insertion tip I diameter of axial distance from a nozzle hole 6.

Comment: This attribute is evident in every drawing of example embodiments of the present invention showing the nozzle insertion tip 1 in combination with the dimensions of said tip I stated in the preceding part of the Description, but is fully graphically evident in FIG 5,6, 12, 14, and 16.

Claim 4. The fuel injector according to any of the preceding claims, wherein its valve spring means 3 is disposed in its entirety within the nozzle insertion tip 1.

Comment: This attribute is evident in FIG 6A, in contrast to the embodiment of FIG 6, where this condition fails to be satisfied by a small margin.

Claim 5. The fuel injector according to any of the preceding claims, wherein the axial component of the distance of a valve closer unit guidance radial support point 27, for aligning and centering the valve closer unit 10 into its valve sealing seat, from the largest diameter ring of sealing contact made by the valve tip section 20 of valve closer unit 10 to its valve seat, is less than the diameter of its valve spring means 3.

Comment: This condition, illustrated in drawing FIG 3, 10, 15, and 17 as the valve closer unit guidance radial support point 27, specifies the stability and accuracy of guidance of the valve precisely into its valve seat. In specifying "a", or any, radial point of support at any given axial location, the point of maximum radial support is included, which is the point axially closest to the target for the precision valve guidance of the valve closer unit 10, more specifically its valve tip section 20, and even more specifically its valve tip congruent seating region 21 where, which is the vaive seat (annular seaiing contact area(s)), which lies as an engineering option anywhere within the valve seat congruent seating region 22, the nomenclature being self-explanatory in the light of the drawings. Of course, the clearance between the polished precision cylindrical bearing surface of the valve lift piston 9 and its similarly precise supporting bore also plays a role in this valve guiding process, which is independent of the present condition for the valve closer unit guidance radial support point 27, which is described in detail near the end of the description for drawing FIG 3. All of the drawings example embodiments of the present invention showing the nozzle insertion tip I satisfy the condition of this claim.

Claim 6. The fuel injector according to any of the preceding claims, wherein the axial component of the distance of a valve closer unit guidance radial support point 27, for aligning and centering the valve closer unit 10 into its valve sealing seat, from the largest diameter ring of sealing contact made by the valve tip section 20 of valve closer unit 10 to its valve seat, is less than the diameter of the nozzle valve lift piston 9.

Comment: The same comment as for Claim 5 applies also to Claim 6. In Drawing FIG 17 the axial position of the ring of sealing contact is ambiguous, and may be chosen by the designer to violate the present claim.

Claim 7. The fuel injector according to any of the preceding claims, wherein, in the closed position of the nozzle valve, the furthest axial component of distance from a point at the tip of the valve tip section 20 to any point on the surface of the valve cylinder chamber 8 is less than two times the diameter of the valve spring means 3.

Comment: This condition is satisfied by all of the drawn example embodiments of the present invention showing the nozzle insertion tip 1, and is an important condition for optimal designs of the present invention, although not an essential condition.

Claim 8. The fuel injector according to any of the preceding claims, wherein the hydraulic cylinder bore of the valve lift piston 9 comprises a valve cylinder chamber 8 bounded in part by the valve lift piston top 11 and said cylinder bore, wherein fluid communication between valve cylinder chamber 8 and a source of unbalancing lower pressure to unbalance a higher pressure in the hydraulic valve lift control chamber 7, bounded by and acting by its pressure on the bottom of the valve lift piston 9, to lift the valve closer unit 10 from its valve seat, is provided by one or more fluid channels opening into the valve cylinder chamber 8 wherein all said fuel channels are defined as being excluded from being part of said cylinder chamber 8; and wherein said hydraulic valve lift control chamber 7 disposed relatively "below" the cylinder chamber 8, which is on the opposite side of the valve lift piston 9 from the cylinder chamber 8 toward the tip of the nozzle insertion tip 1, is similarly served by its separate and distinct one or more fluid channels from a source of high-pressure fuel opening into said hydraulic valve lift control chamber 7, wherein all said fuel channels are defined as excluded from being part of said hydraulic valve lift control chamber 7; and, such that the valve cylinder chamber 8 and the hydraulic valve lift control chamber 7, are entirely contained within the nozzle insertion tip 1.

Comment: This condition is satisfied by all of the drawn example embodiments of the present invention showing the nozzle insertion tip 1.

Claim 9. The fuel injector according to claim 8, wherein said unbalanced higher (valve lifting) pressure in the hydraulic valve lift control chamber 7 is channeled from above the valve cylinder chamber 8 (ie towards the tip of nozzle insertion tip 1) axially past (avoiding) the valve cylinder chamber 8 by a channel, characterized by at least one of: said channel is a lower hydraulic chamber external channel 5A passing axially past the valve cylinder chamber 8 at a radial distance greater than the radius of the valve cylinder chamber 8; said channel is a valve closer unit fluid channel 13 passing axially past (avoiding) the valve cylinder chamber 8 by passing through its piston coupling and guide pipe 18, which in turn passes through its piston coupling and guide pipe bearing 19, which in turn penetrates the cylinder top 12 relatively "upper" boundary of the valve cylinder chamber 8.

Comment: The present claim describes two alternate embodiment approaches, or options for routing high-pressure hydraulic lifting fluid (fuel) past the vaive cylinder chamber 8 to reach the valve lift control chamber 7 from above ("above" meaning the relative to the orientation of the drawing of the fuel injector, and not the installed position of the fuel injector relative to the local surface of our planet). The first option is shown in embodiments of FIG I to 6, and 9. The second option is shown in the embodiments of FIG to 17.

By the way, wherever options are permitted in claims, is the possibility of combining these specified options without restriction in combinations with other features shown in the example options of the present invention which are not specifically illustrated, or in combinations with options not at all shown in the examples of the present invention, without circumventing the claims of the present invention.

Claim 10. The fuel injector according to claims 8 to 9, wherein a fluid channel opening into valve cylinder chamber 8, channeling said source of unbalancing lower pressure to unbalance a higher pressure in its opposing hydraulic valve lift control chamber 7, is characterized by at least one of: the surface of valve cylinder chamber 8 comprises a valve closer unit fluid channel 13 opening into the valve lift piston top 11; the surface of valve cylinder chamber 8 comprises an upper hydraulic chamber external channel 58 opening into the cylinder top 12; Comment: The present claim describes two alternate embodiment approaches, or options for routing low-pressure hydraulic lifting fluid (fuel) into the valve cylinder chamber 8. The first option is shown in embodiments of FIG Ito 6 and 15 to 17. The second option is shown in the embodiments of FIG 10 to 14.

Claim 11. The fuel injector according to claims 8 to 9, wherein a fluid channel opening into valve cylinder chamber 8, channeling said source of unbalancing lower pressure to unbalance a higher pressure in its opposing hydraulic valve lift control chamber 7, is characterized by at least one of: the surface of valve cylinder chamber 8 comprises a flow equalization groove 59 (of any shape, including annular) opening into the valve lift piston top 11, and in fluid communication with a valve closer unit fluid channel 13 at all times, including in the nozzle valve fully open position; the surface of valve cylinder chamber 8 comprises a flow equalization groove 59 (of any shape, including annular) opening into the cylinder top 12, and in fluid communication with a upper hydraulic chamber external channels 58 at all times, including in the nozzle valve fully open position.

Claim 12. The fuel injector according to claims 8 to 11, wherein the valve cylinder chamber 8 is entirely contained within the downstream (tip end) one third (1/3) of the nozzle insertion tip I axial length, or entirely within one nozzle insertion tip I diameter of axial distance from a nozzle hole 6.

Comment: The present claim describes two alternate embodiment approaches, or options, for limiting the axial level of the valve cylinder chamber 8, which are fulfilled by all of the example embodiments of the present invention.

Claim 13. The fuel injector according to claims 8 to 12, wherein the cylinder top 12 surface of the valve cylinder chamber 8 serves as a travel limiting stop for the valve closer unit 10.

Comment: The present claim is fulfilled by all of the example embodiments of the present invention, and is regarded as one of its more significant technical advantages over the state of the art, made possible largely by the small inertial mass of the valve closer unit 10, and the relatively large ratio of the surface area of the cylinder top 12 (about equal to that of the valve lift piston top 11) to the length or the mass of the valve closer unit 10, in comparison with the same measure for a state of the art valve needle. In other words this relatively large valve stopping surface area acts to inhibit valve bounce, or rebound from the stop significantly more than could be possible with the state of the art, as described in the next claim.

is Claim 14. The fuel injector according to claim 13, wherein a valve lift piston 9 of a valve closer unit 10 is lifted by a fuel pressure pulse operative in a hydraulic valve lift control chamber 7 collapsing valve cylinder chamber 8, wherein the cylinder top 12 of the valve cylinder chamber 8 acts as a travel limiting stop comprising a nominally radially symmetric stopping face about the valve lift piston 9 axis (eg planar, spherical, annular corrugated, etc.) substantially congruently contacting the like nominal face of valve lift piston top 11 to limit the stroke of the valve lift piston 9, wherein at a sufficiently small gap between said congruent surfaces, a viscous fluid flow pressure gradient field is established between the approaching or separating congruent surfaces due to the changing restricted surface gap which establishes a retarding and decelerating valve stop approaching force or an adhesive valve separating force upon the valve closer unit 10 which retards the speed of its impact with and possible rebound from said travel limiting stop.

Comment: The same comment as for the previous claim applies here.

Claim 15. The fuel injector according to claim 14, wherein the radially symmetric congruently contacting nominal faces of valve lift piston top 11 and the cylinder top 12 comprise an annular corrugation (increasing the surface area governing the pressure gradient of the viscous fluid flow of the valve cylinder chamber 8 arising at relatively small surface to surface dynamic gap values), the shape and amplitude of which may be adapted to supply the amount of valve closer unit 10 bounce suppression desired, wherein the corrugation cross section may assume for example wave profiles having sinusoidal, triangular, rectangular, mixed, and other shapes.

Comment: This feature is described in detail in the latter part of description for Drawing FIG 1.

Claim 16. The fuel injector according to claims 13 to 15, wherein a flow restriction (eg narrowing) within a fluid channel serving (ie in fluid communication with) the valve cylinder chamber 8, is provided to hydraulically dampen by means of a restricted viscous fluid flow pressure gradient through said fuel flow restriction any degree of unwanted rebound of the valve closer unit 10 from its travel stop, the cylinder top 12.

Comment: This feature is described in the descriptions for Drawings FIG 1 3, and 5.

Claim 17. The fuel injector according to claim 16, wherein said flow restriction is disposed within a valve closer unit fluid channel 13 of the valve closer unit 10.

Comment: This feature is described in the descriptions for Drawings FIG 1, 3, and 5.

Claim 18. The fuel injector according to claims 13 to 17, wherein surface to surface contact of said piston top 11 with said cylinder top 12 sealing around the orifice of a valve closer unit fluid channel opening in a surface of valve cylinder chamber 8 provides a valve sealing surface to surface pressured contact which seals off fluid communication between the switched flow supply channel 44 (via its fluid communication with hydraulic valve lift control chamber 7 and the low pressure supply channel 36, limiting return fuel flow (to the fuel tank, or to a venturi (eg valve closer unit venturi 63), or to any other destination).

Comment: This feature is described in the latter part of description for Drawing FIG 1.

Claim 19. The fuel injector according to any previous claim, wherein an hydraulic channel is connected to a valve closer unit fluid channel 13 of the valve closer unit 10, by means of an axial fluid pipe 17, which passes coaxially through a coil valve spring means 3, which in turn passes coaxially through the insertion tip bore 2 cavity of the nozzle unit 28 and/or its nozzle insertion tip 1, wherein said axial fluid pipe 17 is characterized by any one of: said axial fluid pipe 17 is rigidly attached to the nozzle unit 28 (for example, by a threaded connection), wherein said axial fluid pipe 17 terminates in a dynamic (moving) hydraulic coupling means providing at least leakage resistant isolation from the insertion tip bore 2 cavity (for example by an axial fluid channel coupling slide bearing 16 of the valve closer unit 10, which may be an either "male" or "female" sliding bearing; said axial fluid pipe 17 is rigidly attached to and regarded as a part of the dynamic valve closer unit 10, and makes axially moving (dynamic) hydraulic connection to said hydraulic channel of the fuel injector by dynamic hydraulic coupling means providing at least leakage resistant pressure isolation from the insertion tip bore 2 cavity, such as a sliding bearing connection.

Comment: These features are described in the description for Drawing FIG 1.

Claim 20. The fuel injector according to claim 19, wherein said hydraulic channel connected to a valve closer unit fluid channel 13 is characterized by any one of: said hydraulic channel is the switched flow supply channel 44; said hydraulic channel is the low pressure supply channel 36; said hydraulic channel is a venturi (eg valve closer unit venturi 63).

Comment: Connection of the valve closer unit fluid channel 13 with the switched flow supply channel 44 is presented in conjunction with Drawings 12, 13, 14, 16, and 17.

Connection of valve closer unit fluid channel 13 with the low-pressure supply channel 36 is presented in conjunction with the Drawings FIG 3, 4, 5, and 6. Connection of valve closer unit fluid channel 13 with valve closer unit venturi 63 is presented in conjunction with the Drawings FIG 15, 16, and 17.

Claim 21. The fuel injector according to claim 8 to 20, wherein said axial fluid pipe 17, or its similar truncated axial fluid pipe I 7A (lacking an extended pipe section passing through the coils of the spiral valve spring means 3), is rigidly attached to the nozzle unit 28 (for example, by a threaded connection), wherein it supports either directly, or indirectly by a spacer such as a spring pressure adjustment tube 46, the upper end of the spiral valve spring means 3, and wherein the lower end of the spiral valve spring means 3 is supported by the coil valve spring base plate 14, and wherein the axial fluid pipe 17, or its similar truncated axial fluid pipe I 7A are in fluid communication with the valve closer unit fluid channel 13 passing axially past (avoiding) the valve cylinder chamber 8 by passing through its piston coupling and guide pipe 18, which in turn passes through its piston coupling and guide pipe bearing 19, which in turn penetrates the cylinder top 12 relatively "upper" boundary of the valve cylinder chamber 8 such that said piston coupling and guide pipe 18 emerges along with its valve closer unit fluid channel 13 from said piston coupling and guide pipe bearing 19 on the opposite side from the cylinder top 12, where it joins the coil valve spring base plate 14 50 as to transmit the force of the coil valve spring means 3 to the valve lift piston 9 along with said fluid communication with valve closer unit fluid channel 13.

Comment: Drawings FIG 6, 14, and 16 present the most complete views embodiments of this type.

Claim 22. The fuel injector according to claim 21, wherein said spacer is a spring pressure adjustment tube 46, wherein: said spring pressure adjustment tube 46 has an inner diameter larger than the outer diameter of the shaft of the axial fluid pipe 17 passing through it and through the spiral of the valve spring means 3 supported by spring pressure adjustment tube 46, creating a cylindrical void, the axial fluid pipe spiral flow transfer annular cavity 26B, which is in fluid communication with the nozzle inner spiral fuel channel 25A, and wherein radial apertures, which may be one or more axial fluid pipe inner flow hole(s) I 5B, are provided opening through the spring pressure adjustment tube 46 for fluid communication between the nozzle inner spiral fuel channel 25A and the space radially beyond the outer surface of the spring pressure adjustment tube 46, which is in fluid communication with the nozzle outer spiral fuel channel 24A, and which space is provided fluid communication with a source of hydraulic pressure which is in hydraulic opposition to and isolated from the hydraulic pressure inside the axial fluid pipe 17, wherein one of said two isolated hydraulic pressure spaces carries the flow and pressure of the switched flow supply channel 44, which is channeled to the hydraulic valve lift chamber 7, and wherein the valve cylinder chamber 8 is in fluid communication with the other of said two isolated hydraulic pressure channels, being either the low pressure supply channel 36 or a low pressure supplied by a venturi (eg valve closer unit venturi 63), such that the combination of both isolated hydraulic pressures are able to act separately and respectively at opposite ends of the valve lift piston 9 in their respective opposite hydraulic chambers to lift it.

Comment: Drawings FIG 6 and 14, whose high and low pressure hydraulic channels are reversed in the two embodiments, comprehensively illustrate the present concept, the purpose of which is to insure open channels of flow through the two spiral fuel flow paths the nozzle inner spiral fuel channel 25A, and the nozzle outer spiral fuel channel 24A.

Claim 23. The fuel injector according to claim 22, wherein the lower end of the spiral valve spring means 3 is supported by a coil valve spring base plate 14 which comprises a valve closer spiral flow transfer annular cavity 26A which provides fluid communication facilitating flow from the nozzle inner spiral fuel channel 25A to one or more valve closer inner flow holes 15A, which in turn provide fluid communication between the axially opposite sides of the coil valve spring base plate 14.

Comment: Drawings FIG 3 and 4 illustrate this feature, whose purpose is to reduce resistance to fuel flow through the coil valve spring base plate 14 in comparison to that which exists in the embodiment of FIG 1 and 2.

Claim 24. The fuel injector according to any previous claim, wherein the valve closer unit fluid channel 13 of valve closer unit 10 (which may comprise branches) is at one end (orifice) in fluid communication with the switched flow supply channel 44, and opens at another end through the surface of the valve tip section 20 into fluid communication with the hydraulic valve lift control chamber 7 and its gap volume between the valve seat congruent seating region 22 and the valve tip congruent seating region 21 when the nozzle valve is open, by one or more orifices characterized by at least one of: said orifice opens axially at the tip of the valve tip section 20; said orifice(s) open into the volume of the hydraulic valve lift control chamber 7 wherein the nozzle valve is in its closed position.

Comment: Drawings FIG 13, 13A, and 13B exhibit three variants of this embodiment. See FIG I for the boundaries of the valve tip section 20.

Claim 25. The fuel injector according to claims I to 18 and 24, wherein the valve spring means 3 is an assembly of parallel stacked Belleville springs 62 disposed in the hydraulic valve lift control chamber 7 between the valve lift piston 9 and the valve tip section 20, wherein the latter two parts are connected by a valve to piston shaft 59 comprising a valve closer unit fluid channel 13; and wherein each pair of said Belleville springs 62 is separated by an inner spacer 60 (central spacer) and outer spacer 61 (peripheral spacer); and wherein said assembly (or Belleville valve spring means 3 is mounted upon a valve to piston shaft 59 which joins the valve lift piston (9) above the Belleville valve spring means 3 to the valve tip section 20 below the Belleville valve spring means 3, which is characterized by at least one of: said valve closer unit fluid channel 13 provides fluid communication between the part of the hydraulic valve lift control chamber 7 above the Belleville valve spring means 3, and the part of the hydraulic valve lift control chamber 7 below it, by radial branches ending in orifices, particularly for the transport of fuel past the Belleville valve spring means 3 in the case where the pulsed high pressure fuel of switched flow supply channel 44 for fuel injection is supplied by way of lower hydraulic chamber external channel 5A around the periphery of the valve lift piston 9 to the hydraulic valve lift control chamber 7, wherein the valve lifting (pressure unbalance) and pressure relief need of the valve cylinder chamber 8 is served by the low pressure supply channel 36 connected via the upper hydraulic chamber external channel SB; said valve closer unit fluid channel 13 provides a conduit for fuel injection from the switched flow supply channel 44 as dynamically coupled to the valve closer unit 10 by a piston coupling and guide pipe 18 sliding within a piston coupling and guide pipe bearing (19), and continuing by passing axially through the length of the valve closer unit 10 to its orifice at the tip of the valve tip section 20 supplying the nozzle holes 6 and pulsing the hydraulic valve lift control chamber 7 in conjunction with the valve closer unit venturi 63 within said valve closer unit fluid channel 13 where it passes through the valve lift piston 9, generating its reduced pressure in its venturi branch channels of the valve closer unit fluid channel 13 opening in the surface of the valve lift piston top 11 of the valve cylinder chamber 8, thereby providing its valve lift piston 9 lifting unbalancing pressure; said valve closer unit fluid channel 13 provides a channel for fuel injection from the switched flow supply channel 44 as dynamically coupled to the valve closer unit 10 by a piston coupling and guide pipe (18) sliding within a piston coupling and guide pipe bearing 19, which channel continues by passing axially through the length of the valve closer unit 10 to its orifice at the tip of the valve tip section 20 supplying the nozzle holes 6 and pulsing the hydraulic valve lift control chamber 7, in conjunction with a venturi cylinder 64 unit disposed above the valve cylinder chamber 8 providing valve lift piston 9 lifting low pressure generated by its own valve closer unit venturi 63 disposed within the switched flow supply channel 44, which venturi comprises one or more low pressure generating venturi branch channels, which are upper hydraulic chamber external channel 5B opening into the cylinder top 12 of the valve cylinder chamber 8, thereby providing unbalancing low pressure for lifting its valve lift piston 9.

Comment: Drawings FIG 10 and 11, and FIG 17 graphically describe this claim. In the case of FIG 17 is shown an auxiliary valve closer unit venturi 63 disposed above the nozzle valve as shown in dotted lines, which shows the third of the three optional configurations of the present claim.

Claim 26. For for producing lifting of an hydraulic piston (eg valve lift piston 9) disposed within a fuel injector, is provided a non-return fuel flow means for producing a relatively lower hydraulic pressure in a first hydraulic chamber at a first end (commonly termed "top") of said hydraulic piston (eg valve cylinder chamber 8), wherein a respectively relatively higher lifting pressure in a second hydraulic chamber at the opposite end (commonly termed "bottom") of said hydraulic piston (eg hydraulic valve lift control chamber 7), and in fluid communication with the outlet orifices of the fuel injector's nozzle holes 6 during fuel injections, may be provided directly, or indirectly, from the fuel injection system fuel pressure source to said fuel injector, characterized in that said non-return flow means for producing said relatively lower pressure in said first hydraulic chamber (eg valve cylinder chamber 8), is a venturi (eg valve closer unit venturi 63), either internal or external to said fuel injector, operating by Bernoulli's principle to generate in a venturi branch, or secondary, fluid channel, a relatively reduced pressure in comparison with a relatively higher pressure at the output of its primary flowing fluid pressure drop generating channel, which in turn is a source, direct or indirect, for the pressure in said second (bottom) hydraulic chamber of said hydraulic piston (eg hydraulic valve lift control chamber 7), with the result that the pressure in said first (top) hydraulic chamber (eg valve cylinder chamber 8), is lower than the pressure in said second hydraulic chamber (eg hydraulic valve lift control chamber 7), by virtue of said venturi, causing lifting of said hydraulic piston (eg valve lift piston 9) Comment: Drawings FIG 15, 16, and 17 illustrate example embodiments of the present independent claim.

Claim 27. The fuel injector according to claims I to 25, wherein said source of unbalancing lower pressure to unbalance a higher pressure in the hydraulic valve lift control chamber 7 working on the opposite end from the the valve lift piston top 11 of the valve lift piston 9 to lift the valve closer unit 10 from its valve seat is the low pressure produced by the venturi of claim 26 (ie the valve closer unit venturi 63).

Comment: Drawings FIG 15, 16, and 17 illustrate example embodiments of the present claim, but link them specifically to other claims of the present invention.

Claim 28. The fuel injector according to claim 27, characterized by at least one of said valve closer unit venturi 63 is disposed in the axial (central) bore of the valve closer unit fluid channel 13 within the valve closer unit 10, and the low pressure it produces is channeled by a branch of the valve closer unit fluid channel 13 to the valve cylinder chamber 8; said valve closer unit venturi 63 is disposed in the switched flow supply channel 44, and the low pressure it produces is channeled to the valve cylinder chamber 8; said valve closer unit venturi 63 is disposed in the high pressure supply channel 34 and the low pressure it produces is channeled to the valve cylinder chamber 8; said valve closer unit venturi 63 is disposed in a flow channel of the high pressure fuel supply system for the fuel injector at any point outside of the fuel injector, and the low pressure it produces is supplied to the fuel injector by a fuel return channel connection 31, which in turn is channeled to the valve cylinder chamber 8.

Comment: The present claim expresses the possibility of various optional locations for the valve closer unit venturi. The first named option is illustrated by dashed lines in FIG 17 above the nozzle valve. The second named option is not illustrated, but the high pressure supply channel 34 may be located in FIG 5 and 6, for example. The third option is self explanatory.

Claim 29. A fuel flow controlling (metering or dispensing) piezoelectric operated outward opening metering valve unit 29, suitable for controlling the outflow from a fuel pressure source, such as for example a fuel rail, or a pump, or their combination, etc., for a fuel injector, characterized by a cylindrical valve cavity comprising at one end the mushroom valve 40 valve seat comprising end wall which on its outside separates (isolates) its metering valve fuel inlet chamber 35 into which the mushroom valve 40 opens; and an opposite end wall supporting in its center the coaxial inward-projecting cylinder 53, bored through to support sliding bearing action of the piezoelectric pushrod unit 41 entering from the outside (or alternatively the push rod of an intermediary device such as an hydraulic coupler) to engage and actuate the valve stem of the mushroom valve 40 entering from the inside, wherein the inner and outer mushroom valve 40 stems on both sides of its valve are supported and guided by sliding shaft bearings, the inner stem inside the metering valve unit 29 cylindrical valve cavity in the sliding bearing bore of the coaxial inwardly extending bored coaxial inward-projecting cylinder 53, and the outer stem supported outside the metering valve unit 29 cylindrical valve cavity (on the high pressure supply side of its valve seat) in the outer stem sliding bearing 48 penetrating the opposite surface of the metering valve inlet chamber 35; wherein the mushroom valve 40 is pressed into its valve seat by a coil metering valve spring 57 disposed concentrically within said cylindrical valve cavity of the metering valve unit 29 proximate to its cylindrical wall, its fixed end supported by the valve seat end wall of said cylindrical valve cavity, and its moving end pressed by the spring compressive annular flange of, and its inner coil surface in turn proximate to and coiling around the recessed (flanged hollow cylinder) bearing structure of, the recessed bearing valve spring retainer 50 connected by attachment means at its closed end at a bore accepting the mushroom valve 40 inner stem at an axial position proximate to the end of said coaxial inward-projecting cylinder 53 such that said "recess" of the recessed bearing valve spring retainer 50 proximately encloses said coaxial inward-projecting cylinder 53, optionally by a sliding bearing contact; and wherein metered fuel exits the metering valve unit 29 cylinder through a metering valve outlet port 43.

Comment: The present independent claim protects applications of the present control or metering valve in a broader scope than that of Claim 1, but may optionally include applications falling within the scope of Claim 1. The present control or metering valve is particularly advantageous when applied in combination with the matter of Claim I and all other claims of the present invention, particularly in its unique fast hydraulic pressure transitions between the closed and open state of the valve, characterized as "snapping" actions in claim 34 below. This unique dependency for advantage establishes a necessary unity of invention in the combination of the present claim with the overall invention.

Claim 30. The metering valve unit 29 according to claim 29, wherein a metering valve spiral flow transfer annular cavity 26C recessed below the inner cylindrical surface of the metering valve unit 29 (at a greater radial distance from the cylinder axis than said cylindrical surface) serves the metering valve outlet port 43.

Comment: The metering valve spiral flow transfer annular cavity 26C is best illustrated in FIG 6 and 7, particularly FIG 7, and is described in the Description sections for both Drawings.

Claim 31. The metering valve unit 29 according to claims 29 and 30, wherein the closed-ended cylindrical metering valve unit 29 is created by coaxially assembling the cylinder of coaxial cylinder spring housing 52 within the cylinder of metering valve seat cylinder 56, or vice versa.

Comment: The cylindrical cross sections of the present valve shown best in FIG 7, show clearly the close-fitted assembly of the two cylinders as described in the present claim.

Claim 32. The metering valve unit 29 according to claim 29, wherein by assembling coaxial cylinder spring housing 52 within the metering valve seat cylinder 56 such that the metering valve spiral flow transfer annular cavity 26C is formed as the annular gap remaining where the inner cylindrical wall of the coaxial cylinder spring housing 52 is short of reaching to the bottom of its containing metering valve seat cylinder 56; and wherein the valve outlet port 43 may exit either radially or axially proximate to the valve seat end corner of the metering valve unit 29 cylindrical valve cavity.

Comment: The present claim shows the particularly advantageous formation of the metering valve spiral flow transfer annular cavity 26C by making the inner cylinder sufficiently short in length so as to form said cavity upon assembly.

Claim 33. The fuel injector according to claims I through 28, wherein said pulsed fuel flow controlling (metering or dispensing) means of claim I is the metering valve unit 29 according to claims 29 through 32.

Comment: The present claim combines the independent claim 29 with its specific dependent claims 30 to 32 with the overall invention, namely claims I to 28.

Claim 34. The fuel injector according to claim 33, wherein for a given high pressure supply channel 34 fuel pressure (eg popularly a fuel rail pressure) the piezoelectric lift of the piezoelectric actuator unit 33, the fuel flow limiting throttling effect of the nozzle holes 6, and the valve lift level determined by the valve lift stop cylinder top 12 are coordinated to insure that the closure time of the valve closer unit 10 is reduced by a fuel pressure aided snapping shut of the mushroom valve 40 due to a pressure gradient created by restricted flow through the opening gap between the mushroom valve 40 and its seat as the valve closes, which aids the metering valve spring 57 in closing the mushroom valve 40, a phenomenon always present under all circumstances, but whose relative effect and benefit may be increased by limiting the piezoelectric lift (in most cases being practically the nominal no-load piezoelectric lift due to the relatively negligible metering valve spring 57 force) to a value producing a value of flow through the valve seat substantially at or near the flow limit of the nozzle holes, wherein the nozzle valve lift stop cylinder top 12 is likewise set to stop the opening travel of the valve closer unit 10 to be at or near this same flow limit through the nozzle valve seat, whereby upon retraction of the piezoelectric actuator unit 33 applying the force of the metering valve spring 57 to the mushroom valve 40, a pressure gradient immediately begins to appear and grow through the mushroom valve 40 as it begins to close, aiding its closure by a positive feedback "snap" action of increasing valve closing force due to fuel pressure gradient through the closing mushroom valve 40 gap, which closing force will even reduce the normal no-load closing time of the piezoelectric actuator unit 33, resulting in fast shut off of fuel flow to the nozzle valves 6, resulting in the near immediate collapse of the fuel pressure in the hydraulic valve lift control chamber 7 through the nozzle holes 6 due to the relative incompressibility of the fuel, its limited compressed volume in the switched flow supply channel 44 and within insertion tip bore 2, and the fast closure of the low inertial mass nozzle valve closer unit 10 by its valve spring means 3.

Comment: The present claim exploits the unique advantage for the operation of the present hydraulic fuel injector of the rapid opening and closing speed of an outwardly opening mushroom valve which is operated at low valve lifts which are characteristic of direct piezoelectric actuation, piezoelectric element working strokes being characteristically small, such that the pressure gradient through the small valve lift is present, or nearly present in the valve open position. It is this pressure gradient which causes a snap open and snap shut action of this valve, which is ideal for creating very rapid pressure transitions required for causing the maximum speed of operation for the nozzle valve of Claim 1. And the invention of Claim I is unique in reducing the inertial mass of the nozzle valve below anything offered by the state of the art specifically in order for this nozzle valve to be responsive to maximally fast transitions of its operating hydraulic pressure.

Claim 35. The fuel injector according to any previous claim, wherein the fuel injector is the valve covered orifice (VCO) type.

Claim 36. The fuel injector of claim 35 wherein the VCO nozzle valve incorporates a nozzle hole pressure equalizing groove 23 to improve speed of fuel pressure rise and fall at the nozzle holes 6 on opening and closing the nozzle valve and to improve nozzle hole 6 fuel pressure at low levels of valve closer unit 10 valve lift, permitting reduced levels of valve lift as determined by the valve lift stop cylinder top 12, increasing the fuel injector's injection speed and frequency and decreasing its wear.

Comment: The nozzle hole pressure equalizing groove 23 has been patented by the present author under a FCT application, number FCTJI B2012/051 770, priority date 09 May 2011. The effect of said groove 23 is to substantially increase the speed of pressure change at the nozzle holes 6 due to the opening and closing action of the valve closer unit 10. In other words, for a stop limited, short-stroke, short travel nozzle valve operation as is specified for the present invention, application of said groove 23 would be an ideal enhancement, providing increased power and torque performance, with the advantage of computer controlled ultra fine combustion management by means of an unprecedented number of injection pulses per combustion event, reducing pollution while increasing performance. Also fuel plume penetration is expected to benefit by said groove 23, enabling use of smaller nozzle hole 6 diameter for production of a finer, more efficient spray.

Claim 37. The fuel injector according to any previous claim, wherein attachment of the nozzle insertion tip lower part lB to the nozzle insertion tip upper part IA may be characterized by one or more of: brazing techniques; cold bonding using Surface Activated Bonding (SAB), accomplished for example by atomic surface activation of the surfaces to be joined by ion irradiation, fast atom beam (FAB) irradiation, or hydrogen radical beam irradiation; application of high pressure to the joint while heating the joint.

Comment: This assembly and its bonding options are discussed in the Description for Drawing FIG 3, where it is also illustrated.

Claim 38. The fuel injector according to claims 27 through 28, wherein a valve closer unit venturi 63 is disposed in a valve closer unit fluid channel 13 of valve closer unit 10, characterized by at least one of: downstream in the valve closer unit fluid channel 13 of fluid flow from said venturi 63 is is disposed a branch of said valve closer unit fluid channel 13 establishing fluid communication with the hydraulic valve lift control chamber 7; the portion (segment or section) of the valve closer unit fluid channel 13, from which said connecting branch establishes fluid communication with the hydraulic valve lift control chamber 7, is of diameter which is larger than the inlet diameter of said venturi, for the purpose of increasing the fluid pressure above that of the inlet of said venturi according to Bernoulli's principle; a check valve for limitation of fuel leakage from the nozzle holes 6 is disposed at the outlet orifice of the valve closer unit fluid channel 13 of fluid flow from said venturi 63. which opens at the tip of the valve closer unit 10, which is below (farther toward the nozzle tip of) the inlet orifices of the nozzle holes 6.

Comment: This claim pertains to the embodiments of Drawings FIG 15, FIG ISA, and FIG

15B, and their Description.

Claim 39. The fuel injector according to claims I through 25, and 33 through 38, wherein the valve lift piston 9 is functionally eliminated in terms of its valve closer unit 10 lifting function (it may or may not be physically present, but if it is present, then the hydraulic valve lift control chamber 7 below the piston has free fluid connection to the valve cylinder chamber 8 above the piston, establishing permanent pressure equalization between these two hydraulic chambers), whereby metered switched flow supply channel 44 fuel injection pulses enter into the valve closer unit fluid channel 13 of the piston coupling and guide pipe 18 of the valve closer unit 10, which is connected to the valve tip section 20 of the valve closer unit 10, wherein said valve closer unit fluid channel 13 continues through said valve tip section 20, establishing fluid communication with the trapped volume of fuel below a valve seat ring of valve sealing contact and below the inlet orifices of the nozzle holes 6, such that said injection pulses act upon said valve seat ring of sealing contact, exerting a force of pressure times the ring area in the opening direction of the nozzle valve, which ring diameter exceeds the diameter of the piston coupling and guide pipe bearing 19 and its piston coupling and guide pipe 18 subject to the same injection pulse pressure, but in the closing direction of the nozzle valve, thereby establishing a differential force which opens the nozzle valve, without any return flow being produced in the process.

Comment: This claim pertains to Drawings FIG 18, FIG 19, and FIG 19A, and their

Description.

Claim 40. The fuel injector according to claims 1 through 25, 27, 28, and 33 through 39, wherein two or more metered pressurized fluid (gas or liquid, such as compressed natural gas, liquified petroleum gas, gasoline, diesel fuel, water, etc.) fluid source channels are joined by means of check valves preventing fluid inflow from one fluid source channel into another source channel, providing their flow to a single common receiving fluid channel, the combined switched flow supply channel 70, which conducts said metered fluids to the nozzle valve of said fuel injector for injection into the engine combustion chamber, allowing said fluids to be injected either separately or in any sequence or combination during a combustion process.

Comment: This claim pertains to Drawings FIG 21 and FIG 21A, and their Description.

The present invention is not limited to the descriptions and illustrations of its example embodiments, but embraces other embodiment possibilities within the limits of the claims.

Claims

CLAIMS1. A direct injection inwardly opening,,differential valve type nozzle" based fuel injector, otherwise characterized as operating in a broad sense (allowing for inventive improvement) by the principle of,,nozzle differential ratio, often generically termed,,hydraulic" (also known as a,,mechanical"), depicted for purposes of reference orientation with its nozzle insertion tip (1) tip pointing for purposes of descriptive reference and orientation in the relatively "downward" direction, comprising a nozzle insertion tip (1) (optionally divisible into sections such as nozzle insertion tip upper part IA and nozzle insertion tip lower part I B) extending axially toward the combustion chamber beyond a point of combustion chamber pressure sealing surface contact of said injector with the engine (such as for example an annular shoulder of increased diameter, often sealed by a copper gasket, or tapered pipe threads, etc.) or an intermediary gasket, for direct injection communication penetration of a combustion chamber pressure containment structure (for example, a cylinder head thickness, sometimes comprising cooling elements such as liquid coolant passages or air cooled fins, etc.); wherein a compression valve spring means c3) biases a nozzle valve closer unit (10) comprises an hydraulic valve lift piston (9). and a valve tip (20), which in turn comprises a generally tapering valve tip congruent seating region (21), where in turn is comprised one or more valve sealing contact surfaces establishing valve sealing contact with a valve seat congruent seating region (22); which nozzle valve closer unit (10) is operated (displaced) by hydraulic pressure comprised within a valve lift control chamber (7) controlled by a pulsed fuel flow controlling (metering or dispensing) means, (for example an injection pump, or a valve, connected to a fuel pressure source such as a fuel rail, or a pump, or their combination, etc.) and which may be integral and/or external to the hydraulic fuel injector; characterized by the sliding piston bearing surface (establishing said piston operative pressure differential) of the valve lift piston (9) of said valve closer unit (10) being limited by being confined in its entirety within the nozzle insertion tip (1); and its the inertial mass of the valve closer unit 10 is limited by having its compression valve spring means (3) being limited by being confined, to at least part of its axial length, within the nozzle insertion tip (1).
2. The fuel injector according to claim 1, wherein the entire inertial (moving) mass of valve closer unit (10) is confined to the downstream (tip end) half of the nozzle insertion tip (1) length.
3. The fuel injector according to claim I or 2, wherein the sliding piston bearing surface (establishing said piston operative pressure differential) of the valve lift piston (9) of said valve closer unit (10) being limited by being confined within the downstream (tip end) one third (113) of the nozzle insertion tip (1) axial length, or entirely within one nozzle insertion tip (1) diameter of axial distance from a nozzle hole (6).
4. The fuel injector according to any of the preceding claims, wherein its valve spring means (3) is disposed in its entirety within the nozzle insertion tip (1).
5. The fuel injector according to any of the preceding claims, wherein the axial component of the distance of a valve closer unit guidance radial support point (27), for aligning and centering the valve closer unit (10) into its valve sealing seat, from the largest diameter ring of sealing contact made by the valve tip section (20) of valve closer unit (10) to its valve seat, is less than the diameter of its valve spring means (3).
6. The fuel injector according to any of the preceding claims, wherein the axial component of the distance of a valve closer unit guidance radial support point (27), for aligning and centering the valve closer unit (10) into its valve sealing seat, from the largest diameter ring of sealing contact made by the valve tip section (20) of valve closer unit (10) to its valve seat, is less than the diameter of the nozzle valve lift piston (9).
7. The fuel injector according to any of the preceding claims, wherein, in the closed position of the nozzle valve, the furthest axial component of distance from a point at the tip of the valve tip section (20) to any point on the surface of the valve cylinder chamber (8) is less than two times the diameter of the valve spring means (3).
8. The fuel injector according to any of the preceding claims, wherein the hydraulic cylinder bore of the valve lift piston (9) comprises a valve cylinder chamber (8) bounded in part by the valve lift piston top (11) and said cylinder bore, wherein fluid communication between valve cylinder chamber (8) and a source of unbalancing lower pressure to unbalance a higher pressure in the hydraulic valve lift control chamber (7), bounded by and acting by its pressure on the bottom of the valve lift piston (9), to lift the valve closer unit (10) from its valve seat, is provided by one or more fluid channels opening into the valve cylinder chamber (8) wherein all said fuel channels are defined as being excluded from being part of said cylinder chamber (8); and wherein said hydraulic valve lift control chamber (7) disposed relatively "below" the cylinder chamber (8), which is on the opposite side of the valve lift piston (9) from the cylinder chamber (8) toward the tip of the nozzle insertion tip (1), is similarly served by its separate and distinct one or more fluid channels from a source of high-pressure fuel opening into said hydraulic valve lift control chamber (7), wherein all said fuel channels are defined as excluded from being part of said hydraulic valve lift control chamber (7); and, such that the valve cylinder chamber (8) and the hydraulic valve lift control chamber (7), are entirely contained within the nozzle insertion tip (1).
9. The fuel injector according to claim 8, wherein said unbalanced higher (valve lifting) pressure in the hydraulic valve lift control chamber (7) is channeled from above the valve cylinder chamber (8) (ie towards the tip of nozzle insertion tip (1)) axially past (avoiding) the valve cylinder chamber (8) by a channel, characterized by at least one of: said channel is a lower hydraulic chamber external channel (5A) passing axially past the valve cylinder chamber (8) at a radial distance greater than the radius of the valve cylinder chamber (8); said channel is a valve closer unit fluid channel (13) passing axially past (avoiding) the valve cylinder chamber (8) by passing through its piston coupling and guide pipe (18), which in turn passes through its piston coupling and guide pipe bearing (19), which in turn penetrates the cylinder top (12) relatively "upper" boundary of the valve cylinder chamber (8).
10. The fuel injector according to claims 8 to 9, wherein a fluid channel opening into valve cylinder chamber (8), channeling said source of unbalancing lower pressure to unbalance a higher pressure in its opposing hydraulic valve lift control chamber (7), characterized by at least one of: the surface of valve cylinder chamber (8) comprises a valve closer unit fluid channel (13) opening into the valve lift piston top (11); the surface of valve cylinder chamber (8) comprises a upper hydraulic chamber external channel (5B) opening into the cylinder top (12);
11. The fuel injector according to claims 8 to 9, wherein a fluid channel opening into valve cylinder chamber 8, channeling said source of unbalancing lower pressure to unbalance a higher pressure in its opposing hydraulic valve lift control chamber 7, is characterized by at least one of: the surface of valve cylinder chamber 8 comprises a flow equalization groove 59 (of any shape, including annular) opening into the valve lift piston top 11, and in fluid communication with a valve closer unit fluid channel 13 at all times, including in the nozzle valve fully open position; the surface of valve cylinder chamber 8 comprises a flow equalization groove 59 (of any shape, including annular) opening into the cylinder top 12, and in fluid communication with a upper hydraulic chamber external channels SB at all times, including in the nozzle valve fully open position.
12. The fuel injector according to claims 8 to 11, wherein the valve cylinder chamber (8) is entirely contained within the downstream (tip end) one third (1/3) of the nozzle insertion tip (1) axial length, or entirely within one nozzle insertion tip (1) diameter of axial distance from a nozzle hole (6).
13. The fuel injector according to claims 8 to 12, wherein the cylinder top (12) surface of the valve cylinder chamber (8) serves as a travel limiting stop for the valve closer unit (10).
14. The fuel injector according to claim 13, wherein a valve lift piston (9) of a valve closer unit (10) is lifted by a fuel pressure pulse operative in a hydraulic valve lift control chamber (7) collapsing valve cylinder chamber (8), wherein the cylinder top (12) of the valve cylinder chamber (8) acts as a travel limiting stop comprising a nominally radially symmetric stopping face about the valve lift piston (9) axis (eg planar, spherical, annular corrugated, etc.) substantially congruently contacting the like nominal face of valve lift piston top (11) to limit the stroke of the valve lift piston (9), and wherein said contact collapses substantially all of the volume of the valve cylinder chamber (8) between their nominal congruent surfaces, wherein at a sufficiently small gap between said congruent surfaces, a viscous fluid flow pressure gradient field is established between the approaching or separating congruent surfaces due to the changing restricted surface gap which establishes a retarding and decelerating valve stop approaching force or an adhesive valve separating force upon the low inertial mass valve closer unit (10) which retards the speed of its impact with and possible rebound from said travel limiting stop.
15. The fuel injector according to claim 14, wherein the radially symmetric congruently contacting nominal faces of valve lift piston top (11)and the cylinder top (12) comprise an annular corrugation (increasing the surface area governing the pressure gradient of the viscous fluid flow of the valve cylinder chamber (8) arising at relatively small surface to surface dynamic gap values), the shape and amplitude of which may be adapted to supply the amount of valve closer unit (10) bounce suppression desired, wherein the corrugation cross section may assume for example wave profiles having sinusoidal, triangular, rectangular, mixed, and other shapes.
16. The fuel injector according to claims 13 to 15, wherein a flow restriction (eg narrowing) within a fluid channel serving (ie in fluid communication with) the valve cylinder chamber (8), is provided to hydraulically dampen by means of a restricted viscous fluid flow pressure gradient through said fuel flow restriction any degree of unwanted rebound of the valve closer unit (10) from its travel stop, the cylinder top (12).
17. The fuel injector according to claim 16, wherein said flow restriction is disposed within a valve closer unit fluid channel (13) of the valve closer unit (10).
18. The fuel injector according to claims 13 to 17, wherein surface to surface contact of said piston top (II) with said cylinder top (12) sealing around the orifice of a valve closer unit fluid channel opening in a surface of valve cylinder chamber (8) provides a valve sealing surface to surface pressured contact which seals off fluid communication between the switched flow supply channel (44) (via its fluid communication with hydraulic valve lift control chamber (7) and the low pressure supply channel (36), limiting return fuel flow (to the fuel tank, or to a venturi (eg valve closer unit venturi (63)), or to any other destination).
19. The fuel injector according to any previous claim, wherein an hydraulic channel is connected to a valve closer unit fluid channel (13) of the valve closer unit (10), by means of an axial fluid pipe (17), which passes coaxially through a coil valve spring means (3), which in turn passes coaxially through the insertion tip bore (2) cavity of the nozzle unit (28) and/or its nozzle insertion tip (1), wherein said axial fluid pipe (17) is characterized by any one of: said axial fluid pipe (17) is rigidly attached to the nozzle unit (28) (for example, by a threaded connection), wherein said axial fluid pipe (17) terminates in a dynamic (moving) hydraulic coupling means providing at least eakage resistant isolation from the insertion tip bore (2) cavity (for example by an axial fluid channel coupling slide bearing (16) of the valve closer unit (10), which may be an either "male" or "female" sliding bearing; said axial fluid pipe (17) is rigidly attached to and regarded as a part of the dynamic valve closer unit (10), and makes axially moving (dynamic) hydraulic connection to said hydraulic channel of the fuel injector by dynamic hydraulic coupling means providing at least leakage resistant pressure isolation from the insertion tip bore (2) cavity, such as a sliding bearing connection;
20. The fuel injector according to claim 19, wherein said hydraulic channel connected to a valve closer unit fluid channel (13) is characterized by any one of: said hydraulic channel is the switched flow supply channel (44); said hydraulic channel is the low pressure supply channel (36); said hydraulic channel is a venturi (eg valve closer unit venturi (63)).
21. The fuel injector according to claim 8 to 20, wherein said axial fluid pipe (17), or its similar truncated axial fluid pipe (17A) (lacking an extended pipe section passing through the coils of the spiral valve spring means (3)), is rigidly attached to the nozzle unit (28) (for example, by a threaded connection), wherein it supports either directly, or indirectly by a spacer such as a spring pressure adjustment tube (46), the upper end of the spiral valve spring means (3), and wherein the lower end of the spiral valve spring means (3) is supported by the coil valve spring base plate (14), and wherein the axial fluid pipe (17), or its similar truncated axial fluid pipe (17A) are in fluid communication with the valve closer unit fluid channel (13) passing axially past (avoiding) the valve cylinder chamber (8) by passing through its piston coupling and guide pipe (18), which in turn passes through its piston coupling and guide pipe bearing (19), which in turn penetrates the cylinder top (12) relatively "upper" boundary of the valve cylinder chamber (8) such that said piston coupling and guide pipe (18) emerges along with its valve closer unit fluid channel (13) from said piston coupling and guide pipe bearing (19) on the opposite side from the cylinder top (12), where it joins the coil valve spring base plate (14) so as to transmit the force of the coil valve spring means (3) to the valve lift piston (9) along with said fluid communication with valve closer unit fluid channel (13).
22. The fuel injector according to claim 21, wherein said spacer is a spring pressure adjustment tube (46), wherein: said spring pressure adjustment tube (46) has an inner diameter larger than the outer diameter of the shaft of the axial fluid pipe (17) passing through it and through the spiral of the valve spring means (3) supported by spring pressure adjustment tube (46), creating a cylindrical void, the axial fluid pipe spiral flow transfer annular cavity (26B), which is in fluid communication with the nozzle inner spiral fuel channel (25A), and wherein radial apertures, which may be one or more axial fluid pipe inner flow hole(s) (15B), are provided opening through the spring pressure adjustment tube (46) for fluid communication between the nozzle inner spiral fuel channel (25A) and the space radially beyond the outer surface of the spring pressure adjustment tube (46), which is in fluid communication with the nozzle outer spiral fuel channel (24A), and which space is provided fluid communication with a source of hydraulic pressure which is in hydraulic opposition to and isolated from the hydraulic pressure inside the axial fluid pipe (17), wherein one of said two isolated hydraulic pressure spaces carries the flow and pressure of the switched flow supply channel (44), which is channeled to the hydraulic valve lift chamber (7), and wherein the valve cylinder chamber (8) is in fluid communication with the other of said two isolated hydraulic pressure channels, being either the low pressure supply channel (36) or a low pressure supplied by a venturi (eg valve closer unit venturi (63)), such that the combination of both isolated hydraulic pressures are able to act separately and respectively at opposite ends of the valve lift piston (9) in their respective opposite hydraulic chambers to lift it.
23. The fuel injector according to claim 22, wherein the lower end of the spiral valve spring means (3) is supported by a coil valve spring base plate (14) which comprises a valve closer spiral flow transfer annular cavity (26A) which provides fluid communication facilitating flow from the nozzle inner spiral fuel channel (25A) to one or more valve closer inner flow holes (1 5A), which in turn provide fluid communication between the axially opposite sides of the coil valve spring base plate (14).io
24. The fuel injector according to any previous claim, wherein the valve closer unit fluid channel (13) of valve closer unit (10) (which may comprise branches) is at one end (orifice) in fluid communication with the switched flow supply channel (44), and opens at another end through the surface of the valve tip section (20) into fluid communication with the hydraulic valve lift control chamber (7) and its gap volume between the valve seat congruent seating region (22) and the valve tip congruent seating region (21) when the nozzle valve is open, by one or more orifices characterized by at least one of: said orifice opens axially at the tip of the valve tip section (20); said orifice(s) open into the volume of the hydraulic valve lift control chamber (7) wherein the nozzle valve is in its closed position.
25. The fuel injector according to claims Ito 18 and 24, wherein the valve spring means (3) is an assembly of parallel stacked Belleville springs (62) disposed in the hydraulic valve lift control chamber (7) between the valve lift piston (9) and the valve tip section (20), wherein the latter two parts are connected by a valve to piston shaft (59) comprising a valve closer unit fluid channel (13); and wherein each pair of said Belleville springs (62) is separated by an inner spacer (60) (central spacer) and outer spacer (61) (peripheral spacer); and wherein said assembly (or Belleville valve spring means (3) is mounted upon a valve to piston shaft (59) which joins the valve lift piston (9) above the Belleville valve spring means (3) to the valve tip section (20) below the Belleville valve spring means (3), which is characterized by at least one of: said valve closer unit fluid channel (13) provides fluid communication between the part of the hydraulic valve lift control chamber (7) above the Belleville valve spring means (3), and the part of the hydraulic valve lift control chamber (7) below it, by radial branches ending in orifices, particularly for the transport of fuel past the Belleville valve spring means (3) in the case where the pulsed high pressure fuel of switched flow supply channel (44) for fuel injection is supplied by way of lower hydraulic chamber external channel (5A) around the periphery of the valve lift piston (9) to the hydraulic valve lift control chamber (7), wherein the valve lifting (pressure unbalance) and pressure relief need of the valve cylinder chamber (8) is served by the low pressure supply channel (36) connected via the upper hydraulic chamber external channel (5B); said valve closer unit fluid channel (13) provides a conduit for fuel injection from the switched flow supply channel (44) as dynamically coupled to the valve closer unit (10) by a piston coupling and guide pipe (18) sliding within a piston coupling and guide pipe bearing (19), and continuing by passing axially through the length of the valve closer unit (10) to its orifice at the tip of the valve tip section (20) supplying the nozzle holes (6) and pulsing the hydraulic valve lift control chamber (7) in conjunction with the valve closer unit venturi (63) within said valve closer unit fluid channel (13) where it passes through the valve lift piston (9), generating its reduced pressure in its venturi branch channels of the valve closer unit fluid channel (13) opening in the surface of the valve lift piston top (11) of the valve cylinder chamber (8), thereby providing its valve lift piston (9) lifting unbalancing pressure; said valve closer unit fluid channel (13) provides a channel for fuel injection from the switched flow supply channel (44) as dynamically coupled to the valve closer unit (10) by a piston coupling and guide pipe (18) sliding within a piston coupling and guide pipe bearing (19), which channel continues by passing axially through the length of the valve closer unit (10) to its orifice at the tip of the valve tip section (20) supplying the nozzle holes (6) and pulsing the hydraulic valve lift control chamber (7), in conjunction with a venturi cylinder (64) unit disposed above the valve cylinder chamber (8) providing valve lift piston (9) lifting low pressure generated by its own valve closer unit venturi (63) disposed within the switched flow supply channel (44), which venturi comprises one or more low pressure generating venturi branch channels, which are upper hydraulic chamber external channel (5B) opening into the cylinder top (12) of the valve cylinder chamber (8), thereby providing unbalancing low pressure for lifting its valve lift piston (9).
26. For for producing lifting of an hydraulic piston (eg valve lift piston (9)) disposed within a fuel injector, is provided a non-return fuel flow means for producing a relatively lower hydraulic pressure in a first hydraulic chamber at a first end (commonly termed "top") of said hydraulic piston (eg valve cylinder chamber (8)), wherein a respectively relatively higher lifting pressure in a second hydraulic chamber at the opposite end (commonly termed "bottom") of said hydraulic piston (eg hydraulic valve lift control chamber (7)), and in fluid communication with the outlet orifices of the fuel injector's nozzle holes (6) during fuel injections, may be provided directly, or indirectly, from the fuel injection system fuel pressure source to said fuel injector, characterized in that said non-return flow means for producing said relatively lower pressure in said first hydraulic chamber (eg valve cylinder chamber (8)), is a venturi (eg valve closer unit venturi (63)), either internal or external to said fuel injector, operating by Bernoulli's principle to generate in a venturi branch, or secondary, fluid channel, a relatively reduced pressure in comparison with a relatively higher pressure at the output of its primary flowing fluid pressure drop generating channel, which in turn is a source, direct or indirect, for the pressure in said second (bottom) hydraulic chamber of said hydraulic piston (eg hydraulic valve lift control chamber (7)), with the result that the pressure in said first (top) hydraulic chamber (eg valve cylinder chamber (8)), is lower than the pressure in said second hydraulic chamber (eg hydraulic valve lift control chamber (7)), by virtue of said venturi, causing lifting of said hydraulic piston (eg valve lift piston (9))
27. The fuel injector according to claims I to 25, wherein said source of unbalancing lower pressure to unbalance a higher pressure in the hydraulic valve lift control chamber (7) working on the opposite end from the the valve lift piston top (11)of the valve lift piston (9) to lift the valve closer unit (10) from its valve seat is the low pressure produced by the venturi of claim 26 (ie the valve closer unit venturi (63)).
28. The fuel injector according to claim 27, characterized by at least one of said valve closer unit venturi (63)is disposed in the axiaL (central) bore of the valve closer unit fluid channel (13) within the valve closer unit (10), and the low pressure it produces is channeled by a branch of the valve closer unit fluid channel (13) to the valve cylinder chamber (8); said valve closer unit venturi (63) is disposed in the switched flow supply channel (44), and the low pressure it produces is channeled to the valve cylinder chamber (8); said valve closer unit venturi (63)is disposed in the high pressure supply channel (34) and the low pressure it produces is channeled to the valve cylinder chamber (8); said valve closer unit venturi (63) is disposed in a flow channel of the high pressure fuel supply system for the fuel injector at any point outside of the fuel injector, and the low pressure it produces is supplied to the fuel injector by a fuel return channel connection (31), which in turn is channeled to the valve cylinder chamber (8).
29. A fuel flow controlling (metering or dispensing) piezoelectric operated outward opening metering valve unit (29), suitable for controlling the outflow from a fuel pressure source, such as for example a fuel rail, or a pump, or their combination, etc., for a fuel injector, characterized by a cylindrical valve cavity comprising at one end the mushroom valve (40) valve seat comprising end wall which on its outside separates (isolates) its metering valve fuel inlet chamber (35) into which the mushroom valve (40) opens; and an opposite end wall supporting in its center the coaxial inward-projecting cylinder (53), bored through to support sliding bearing action of the piezoelectric pushrod unit (41) entering from the outside (or alternatively the push rod of an intermediary device such as an hydraulic coupler) to engage and actuate the valve stem of the mushroom valve (40) entering from the inside, wherein the inner and outer mushroom valve (40) stems on both sides of its valve are supported and guided by sliding shaft bearings, the inner stem inside the metering valve unit (29) cylindrical valve cavity in the sliding bearing bore of the coaxial inwardly extending bored coaxial inward-projecting cylinder (53), and the outer stem supported outside the metering valve unit (29) cylindrical valve cavity (on the high pressure supply side of its valve seat) in the outer stem sliding bearing (48) penetrating the opposite surface of the metering valve inlet chamber (35); wherein the mushroom valve (40) is pressed into its valve seat by a coil metering valve spring (57) disposed concentrically within said cylindrical valve cavity of the metering valve unit (29) proximate to its cylindrical wall, its fixed end supported by the valve seat end wall of said cylindrical valve cavity, and its moving end pressed by the spring compressive annular flange of, and its inner coil surface in turn proximate to and coiling around the recessed (flanged hollow cylinder) bearing structure of, the recessed bearing valve spring retainer (50) connected by attachment means at its closed end at a bore accepting the mushroom valve (40) inner stem at an axial position proximate to the end of said coaxial inward-projecting cylinder (53) such that said "recess" of the recessed bearing valve spring retainer (50) proximately encloses said coaxial inward-projecting cylinder (53), optionally by a sliding bearing contact; and wherein metered fuel exits the metering valve unit (29) cylinder through a metering valve outlet port (43).
30. The metering valve unit (29) according to claim 29, wherein a metering valve spiral flow transfer annular cavity (26C) recessed below the inner cylindrical surface of the metering valve unit (29) (at a greater radial distance from the cylinder axis than said cylindrical surface) serves the metering valve outlet port (43).
31. The metering valve unit (29) according to claims 29 and 30, wherein the closed-ended cylindrical metering valve unit (29) is created by coaxially assembling the cylinder of coaxial cylinder spring housing (52) within the cylinder of metering valve seat cylinder (56), or vice versa.
32. The metering valve unit (29) according to claim 29, wherein by assembling coaxial cylinder spring housing (52) within the metering valve seat cylinder (56) such that the metering valve spiral flow transfer annular cavity (26C) is formed as the annular gap remaining where the inner cylindrical wall of the coaxial cylinder spring housing (52) is short of reaching to the bottom of its containing metering valve seat cylinder (56); and wherein the valve outlet port (43) may exit either radially or axially proximate to the valve seat end corner of the metering valve unit (29) cylindrical valve cavity.is
33. The fuel injector according to claims I through 28, wherein said pulsed fuel flow controlling (metering or dispensing) means of claim I is the metering valve unit (29) according to claims 29 through 32.
34. The fuel injector according to claim 33, wherein for a given high pressure supply channel (34) fuel pressure (eg popularly a fuel rail pressure) the piezoelectric lift of the piezoelectric actuator unit (33), the fuel flow limiting throttling effect of the nozzle holes (6), and the valve lift level determined by the valve lift stop cylinder top (12) are coordinated to insure that the closure time of the valve closer unit (10) is reduced by a fuel pressure aided snapping shut of the mushroom valve (40) due to a pressure gradient created by restricted flow through the opening gap between the mushroom valve (40) and its seat as the valve closes, which aids the metering valve spring (57) in closing the mushroom valve (40), a phenomenon always present under all circumstances, but whose relative effect and benefit may be increased by limiting the piezoelectric lift (in most cases being practically the nominal no-load piezoelectric lift due to the relatively negligible metering valve spring (57) force) to a value producing a value of flow through the valve seat substantially at or near the flow limit of the nozzle holes, wherein the nozzle valve lift stop cylinder top (12) is likewise set to stop the opening travel of the valve closer unit (10) to be at or near this same flow limit through the nozzle valve seat, whereby upon retraction of the piezoelectric actuator unit (33) applying the force of the metering valve spring (57) to the mushroom valve (40), a pressure gradient immediately begins to appear and grow through the mushroom valve (40) as it begins to close, aiding its closure by a positive feedback "snap" action of increasing valve closing force due to fuel pressure gradient through the closing mushroom valve (40) gap, which closing force will even reduce the normal no-load closing time of the piezoelectric actuator unit (33), resulting in fast shut off of fuel flow to the nozzle valves (6), resulting in the near immediate collapse of the fuel pressure in the hydraulic valve lift control chamber (7) through the nozzle holes (6) due to the relative incompressibility of the fuel, its limited compressed volume in the switched flow supply channei (44) and within insertion tip bore (2), and the fast closure of the low inertial mass nozzle valve closer unit (10) by its valve spring means (3).
35. The fuel injector according to any previous claim, wherein the fuel injector is the valve covered orifice (VCO) type.
36. The fuel injector of claim 35 wherein the VCO nozzle valve incorporates a nozzle hole pressure equalizing groove (23) to improve speed of fuel pressure rise and fall at the nozzle holes (6) on opening and closing the nozzle valve and to improve nozzle hole (6) fuel pressure at low levels of valve closer unit (10) valve lift, permitting reduced levels of valve lift as determined by the valve lift stop cylinder top (12), increasing the fuel injector's injection speed and frequency and decreasing its wear.
37. The fuel injector according to any previous claim, wherein attachment of the nozzle insertion tip lower part (1 B) to the nozzle insertion tip upper part (IA) may be characterized by one or more of: brazing techniques; cold bonding using Surface Activated Bonding (SAB), accomplished for example by atomic surface activation of the surfaces to be joined by ion irradiation, fast atom beam (FAB) irradiation, or hydrogen radical beam irradiation; application of high pressure to the joint while heating the joint.38 The fuel injector according to claims 27 through 28 wherein a valve closer unit venturi (63) is disposed in a valve closer unit fluid channel (13) of valve closer unit (10) characterized by at least one of downstrearii in the valve closer unit fluid channel (13) of fluid flow from said venturi (63) is disposed a branch of said valve closer unit fluid channel (13) establishing fluid communication with the hydraulic valve lift control chamber (7); the portion (segment or section) of the valve closer unit fluid channel (13), from which said connecting branch establishes fluid communication with the hydraulic valve lift control chamber 7), is of dianieter which is larger than the inlet diameter of said venturi, for the purpose of increasing the fluid pressure above that of the inlet of said venturi according to Bernoulli's principle: a check valve for limitation of fuel leakage from the nozzle holes (6) is disposed at the outlet orifice of the valve closer unit fluid channel (13) of fluid flow from said venturi (63), which opens at the tip of the valve closer unit (10), which is below (farther toward the nozzle tip of) the inlet orifices of the nozzle holes (6).39. The fuel injector according to claims I through 25, and 33 through
38. wherein the valve lift piston (9) is functionally eliminated in terms of its valve closer unit (10) lifting function (it may or may not be physically present, but if it is present, then the hydraulic valve lift control chamber 7) below the piston has free fluid connection to the valve cylinder chamber (8) above the piston, establishing permanent pressure equalization between these two hydraulic chambers), whereby metered switched flow supply channel (44) fuel injection pulses enter into the valve closer unit fluid channel (13) of the piston coupling and guide pipe (18) of the valve closer unit (10), which is connected to the valve tip section (20) of the valve closer unit (10), wherein said valve closer unit fluid channel (13) continues through said valve tip section (20), establishing fluid communication with the trapped volume of fuel below a valve seat ring of valve sealing contact and below the inlet orifices of the nozzle holes (6), such that said injection pulses act upon said valve seat ring of sealing contact, exerting a force of pressure times the ring area in the opening direction of the nozzle valve, which ring diameter exceeds the diameter of the piston coupling and guide pipe bearing (19) and its piston coupling and guide pipe (18) subject to the same injection pulse pressure, but in the closing direction of the nozzle valve, thereby establishing a differential force which opens the nozzle valve, without any return flow being produced in the process.40. The fuel injector according to claims I through 25, 27, 28, and 33 through 39, wherein two or more metered pressurized fluid (gas or liquid, such as compressed natural gas, liquified petroleu in gas, gasoline, diesel fuel, water, etc.) fluid source channels are joined by means of check valves preventing fluid inflow from one fluid source channel into another source channel providing their flow to a single common receiving fluid channel the combined switched flow supply channel (70), which conducts said metered fluids to the nozzle valve of said fuel injector for injection into the engine combustion chamber allowing said fluids to be injected either separately or in any sequence or combination dunng a combustion process