Within this patent the terminology used is as follows:- The term'super-charger'is used, for convenience herein, to embrace any form of fluid, in particular gas, flow-booster, assister, enhancer, promoter-whether with, or without, fluid compression (internally of the device).

A prime super-charger use is for (rein) forced or ram flow of a combustion gas component-such as air-into the intake of an i. c. engine.

Broadly, the objective is to introduce a greater charge mass into a combustion chamber upon, or rather preparatory to, each combustion cycle-than the engine could otherwise achieve on its own (unassisted) account.

In common usage, the term super-charger implies a mechanical drive -typically directly, or indirectly, from an engine output shaft, or alternatively by an independent ancillary (say, electric) motor.

One particular category of indirect drive, designated turbo- charging (strictly-speaking'turbo-superchargingl, uses surplus (and so otherwise waste) energy from engine exhaust gas flow, to drive a turbine, in turn coupled to a compressor.

Generally-absent (airflow) boosting or enhancement measures, such as superchargers, to re-inforce, or supplement the intake air flow-i. c. engines are reliant upon their inherent'suction', or pressure differential, pumping ability.

Engine output reflects the pressure and density of the combustion charge-and, absent supplementary boost measures, such as achieved by a super-charger in an intake path-is constrained by the engine'self-charge'ability.

Engines can thus benefit from super-charging, primarily because of the additional power that can be produced-for a given engine size (capacity) and speed-by additional airflow through the engine.

Set against this, has to be offset the penalties of mechanical complexity, weight, space and cost.

Supercharging is known in single and multi-stages.

Any simplification, or efficiency enhancement, of the supercharging system is potentially advantageous-provided it is achieved without sacrificing reliab lity or durability.

Of the various types of super-charger applied commercially, only the so-called'turbo-super-charger'-commonly abbreviated to 'turbocharger', has proved sufficiently reliable and efficient for routine application to commercial vehicle diesel engines.

For commercial vehicles, where fuel consumption is a critical cost of operation, the inherent ability of a turbocharger to'self- regulate'-that is to provide extra air only as required- significantly improves engine fuel consumption, over other super- charger types.

Broad categories of super-charger configuration include the so- called: 'Rootes'blower-with straight, or somewhat helical, fluted intermeshing rotor pairs, within a complementary profiled outer housing or casing; no internal pressure gain arises, although downstream pressure can be higher than blower intake pressure, if the engine's natural air flow is less than that of the blower.

Centrifugal compressor, with dynamic compression; Screw compressor, similar to a Rootes blower, but with a complementary angular twist introduced into intermeshing, complementary profile male/female rotors-and ports suitably dimensioned to achieve flow compression, as the fluid (ie air) passes therebetween and alongside the rotor axes.

The process of raising the pressure involves some temperature rise, and screw compressors can impart less heat than Rootes type blowers-and so deliver a denser air mass and concomitantly greater charge per unit volume.

Thus, the effect of a super-charger upon engine performance, can be expressed as increased'power density'-due to: * combustion chamber charge pressure; * and density of the combustion mixture.

Rootes super-chargers were popular in early automobile gasoline engine development.

Whilst they create no internal compression, reliance is placed upon driving or forcing a predetermined volume of air-namely that ingested into the super-charger working chamber (s)-into a smaller output volume-namely that of the engine cylinder (s).

The overall objective is to increase both charge pressure and density.

However, considerable back-flow takes place-making the process inefficient.

In a conventional screw compressor, generally helical, or rather helical-fluted, paired rotors intermesh, in synchronism, by external inter-coupled gearing.

The working fluid-ie air-is entrapped between the rotors and the surrounding housing.

The rotors rotate, in opposite directions, drawing the working fluid, or air, longitudinally therebetween, and alongside the rotor axes-and trapping it in closed chambers, which gradually reduce in volume.

Thus the working fluid, or air, becomes progressively more compressed.

Overall, between an inlet port, typically at one rotor end, and an outlet port, typically in the housing circumference, at an opposite end from the inlet, an internal pressure differential- specifically a pressure rise or gain-is created, together with an enhanced (air) through-flow.

Rotors are enmeshed with minimal clearance between instantaneous 'contact points', or points of closest proximity, allowing for any marginal working clearance between them.

Any rotor contact is desirably through an intervening lubricant film-in order to minimise leakage between the rotors.

In any event leakage may prove less of a factor with increased rotor speed.

The'internal compression'attendant a conventional screw compressor can increase efficiency compared with a Rootes type blower.

For this reason a compressor for the typical industrial application with significant pressure ratio (c. 6: 1 is common) is now of screw type; Rootes and piston compressors now have a very small market share.

This is because the cost of the energy of operation generally dominates the full life costing of the equipment and increased efficiency is thus very important in the economic equation.

For lower pressures non-positive displacement devices are favoured although there is still some market for Rootes blowers; screw compressors are not generally used because of their high initial cost (dominated by the cost of manufacturing the complex rotor geometry).

For high pressure ratios (greater than approximately 4: 1) the temperature of the compressed air is sufficient to degrade any mineral oil mist or other organic lubricant present.

Deposits of the degraded substances would be left on the internal surfaces of the blower causing fouling and eventually either grossly inefficient operation or component failure.

Compressors for high pressure ratios either have to be run oil- free to ensure deposits do not accumulate, or be"flooded"with oil (or other fluid) where one of the primary duties of the oil (or other fluid) is to cool the compressed gas (air) to prevent the above mentioned degradation and fouling.

Internal cooling by oil (or other fluid) has the disadvantage that additional systems are required for adding the correct quantity of oil and removing it again after compression.

"Gearless"screw compressors (wherein there is no separate synchronising or timing gear between the rotors) are known, but these are used with oil flooding to ensure that the contact between the rotors does not cause intolerable wear.

Even so great care must be taken in design of the rotor profiles to ensure successful operation.

A continuously driven, mechanically coupled (conventional) screw compressor shares a disadvantage of a Rootes blower-namely, of compressing the charge (through the engine)-regardless of whether or not increased charge pressure is required.

Constructionally, industrial screw compressors typically have steel rotors and timing gears, and cast iron housings.

The resulting heavy units would offset much of the improvement in 'power density', that would otherwise be expected to occur-if applied to engines as super-chargers.

More recently, screw compressors have been produced that use light alloy rotors and housings, but these have not been widely used in automotive applications.

Some aspects of the present invention address forced induction systems, suitable for aircraft engines, employing variants of screw compressor super-charger technology.

Thus, it will be seen that embodiments of the invention embrace lower internal pressure ratios than a conventional screw compressor-even to the extent, in certain circumstances, of no, or at least minimal, significant internal compression.

Thus, whilst the term'screw compressor'is convenient, in at least reflecting the origins in the art, the broader, albeit informal, terms'screw displacer'or'screw blower'are employed later in the present disclosure, and in the appended claims, in order to avoid undue limitation or confusion from the term 'compressor'when in practice no such compression may arise.

According to one aspect of the present invention, a screw compressor, -such as for use in conjunction with an i. c. engine incorporates complementary intermeshing, but otherwise non-mechanically coupled, helical rotors, with an active (driving) rotor, and a passive (driven) rotor, with intake and outlet ports configured to achieve an internal'pressure ratio', in'normal'operation, of between unity (representing no significant compression), and some desired valued-for example, of less than 3: 1 Preferably, the internal pressure ratio is less than 1.5: 1 in 'normal'operation.

Given suitable rotor materials and coatings, along with an intervening working fluid-in this case air, the intermeshing of rotor profiles effectively transfers drive'internally', from one rotor to another.

This obviates the supplementary external timing gears, gear housing, and attendant lubrication system of a conventional screw compressor.

Such constructional simplification is without loss of critical operational features.

Moreover very low torque reaction arises from internal air pressures.

For design economy, it is envisaged that suitable rotor configurations could be adapted from known (mechanical gear inter- coupled) screw compressors.

The rotor profiles disclosed in UK Patent Application 9610289.2 are particularly suitable for operation without external synchronising means. UK Patent Application 9923522.8 discusses a specific design (& profile) that enhances this still further.

The rotors could employ low density, and so light-weight, metal alloys, synthetic plastics or composites.

Similarly, the super-charger housings could feature low-density, and so light-weight, metal alloys, synthetic plastics, or composites.

The resulting light rotors and housings reduce the weight of the unit considerably.

For aircraft use, this weight reduction makes a screw compressor viable as a supercharger.

Many low-density materials also have the benefit of corrosion and oxidation resistance-obviating the need for rotor and housing corrosion protection.

If made of carbon or alloy steel and cast iron, rotors and housings would need special provision to prevent corrosion.

According to another aspect of the present invention, a forced-induction apparatus, for an i. c. engine, incorporates a screw-compressor, according to one aspect of the invention, as a'primary', or first-stage, super-charger.

A multi-stage screw-compressor variant could be feasible-with successive stages in tandem-either as discrete devices or integrated within a common housing.

Thus, a secondary'turbo-super-charger' (or simply turbo-charger) can be located upstream (as a pre-boost), or downstream (as a post-boost) of a primary supercharger.

Overall, the forced-induction system may incorporate an inter- cooler and/or an after-cooler.

Some embodiments of the invention are applicable to two-stroke engines.

A two-stoke diesel engine would be a case in point.

Generally, a two-stroke engine requires a positive pressure gradient across it at all times, in order to ensure: * adequate exhaust product scavenge; and * adequate cylinder fill with fresh charge.

A high efficiency turbo-charger can provide this pressure gradient over part of an engine's operating range, but generally cannot provide it at part load, or for starting. Since turbo-chargers do not have the characteristics required of a super-charger such as minimal lag characteristics-in order to ensure reliable two-stroke engine operation, turbo-chargers have not generally been adopted hitherto on two-stroke engines, as a sole means of supercharging, Nevertheless, auxiliary blowers-usually of centrifugal type- and driven electrically, are in widespread use for marine two- stroke engines.

This approach is not readily applicable to automotive engines- and for aircraft use would increase risks unacceptably.

Thus, an electrical failure causing the engine to stop would be unwelcome in an automotive application-but could prove more serious for an aeroplane.

Moreover, fouling of the turbocharger, or at the extreme complete failure, would also cause the engine to stop, because the electric blower would not be able to provide the increased pressure needed under these conditions.

A mechanically-driven radial compressor could be used, but, in order to provide sufficient pressure at starting, it would need to be very large, or be rotated at many times crankshaft speed.

Consequently, it would produce a high (internal)'pressure-ratio', at full engine speed resulting in engine'over-boost'. A clutch could be used to disengage the radial compressor, when not required-but this would add complexity back into the system.

Adoption of a positive-displacement,'gearless'or'uncoupled', screw compressor as a (primary) super-charger-downstream of the compressor of a secondary turbocharger, according to an aspect of the invention-overcomes such problems of conventional stand- alone turbo-chargers.

According to one aspect of the invention, a'gearless'screw super-charger, is envisaged as operational with either unity, or a relatively low pressure ratio, compared with conventional geared compressors.

'Normally', in variants of the invention, the relevant pressure ratio is less then 3: 1- and often, for aircraft engines, less than 1.5: 1.

In this latter case there would be little load on the rotors of the super-charger-at any condition.

Thus, the risk associated with using a'gearless'super-charger configuration is perceived as very low.

A super-charger configured according to the invention would be capable of operation at higher pressure ratios, for short periods.

Thus, for example, in the event of a system failure, such as fouling of the turbocharger-that resulted in a higher pressure gradient being needed to force the air through the engine, a (gearless) super-charger so configured according to the invention could meet the temporary demand.

Otherwise, the (gearless) super-charger is generally only operating at a low pressure ratio-with relatively little power consumption.

Thus, the reduction in (compressor) efficiency, when operated away from its peak, would have little contributory effect upon the engine's overall efficiency.

Again, since the internal pressures are low, the rotors and housings can be produced as light-weight components, using suitable low-density alloy, plastic or composite, with adequate strength and stiffness.

Clearances are also of lesser importance at low pressure ratios- so the components can be made less accurately and at lower cost.

Light weight also results in low inertia-and so reduced loads associated with engine acceleration and torsional vibration.

The low pressure-ratio and modest power consumption also mean that a simple'low-power', or generally insubstantial, drive will be adequate.

A forced-induction apparatus according to an aspect of the invention can be used with a two-stroke diesel aircraft engine.

By placing a'gearless'screw compressor super-charger, according to one aspect of the invention, downstream of the turbocharger compressor and upstream of the engine, the number of components critical to the engine's continued operation-and hence the risks of engine failure-are reduced to the minimum.

In'normal'operation, the turbocharger thus positioned will provide additional airflow hence increasing the engine power, as designed.

The power available with a failed turbocharger-and with the screw super-charger alone providing the pressure gradient across the engine that is needed for continued engine operation-will probably be adequate for continued flight.

The turbocharger may be of any conventional type-such as radial compressor and turbine; radial compressor with axial turbine; radial compressor and mixed flow turbine; etc., available as proprietary products, such as, for example Garret/Allied Signal, Holset, KKK, Mitsubishi, and Brown Boveri.

Careful matching of the engine, turbocharger and super-charger will be required to optimise system operation and performance, under all conditions. According to a further aspect of the invention, a forced-induction apparatus, for an i. c. engine, incorporates a super-charger, configured as a'gearless', low, minimal or zero, internal pressure ratio (or gain), screw-compressor, together with a turbo- charger.

Operation of the forced induction apparatus could be adjusted, should the turbocharger's output be greater than that which requires a super-charger to used for optimum engine operation.

Thus, for example, a one-way (eg sprag type) clutch, might be incorporated into the super-charger drive.

Such a clutch could allow the super-charger rotors to rotate faster than they would have done, if driven by the engine via a solid coupling.

In these circumstances, the super-charger would to some extent operate as a turbine-'expanding'the charge, and using the power thus produced to increase its speed of rotation.

A one-way clutch would also provide some degree of de-coupling of the supercharger, from torsional crankshaft resonances.

One or more by-passes may be provided, from the super-charger inlet to the outlet. These may include valves, that allow charge air to flow only from inlet to outlet-if the inlet pressure is higher than the outlet pressure, by more than a predetermined amount.

This would also improve the operation of the forced induction apparatus-should the turbocharger's output be greater than that which would (conventionally) require a super-charger to be used for optimum engine operation.

Such valves and by-passes may also advantageously be arranged to allow flow, from the outlet, back to the inlet of the super- charger, or to the inlet of the turbo-super-charger, or to atmosphere (advantageously via the engine exhaust system), if the pressures, or pressure differences, rise above other predetermined amounts.

The turbocharger may also include a turbine by-pass, or wastegate -which would also help to guard against excessive engine boost.

In practice, the forced induction apparatus may comprise more than one supercharger and/or turbocharger-in series and/or parallel.

The apparatus may also include inter-coolers and/or after-coolers, as required, between various components.

Several stages of compression, in series, can be used, in order to give still further increase in'charge density'.

Compressors, or blowers, in parallel may be used to improve packaging, or to allow de-clutching of individual compressors/ blowers-as dictated by operational requirements.

There may also be certain circumstances in which a turbocharger compressor could be located downstream (rather than upstream) of a screw compressor.

There now follows a description of some particular embodiments of the invention, by way of example only, with reference to the accompanying schematic and diagrammatic and schematic drawings, in which: Figure 1 shows an overall layout of an i. c. engine and air intake system, with provision for super-charger boost in a (combustion air) intake or induction path; Figure 2 shows a part-sectioned, part cut-away, perspective view of a'gearless'screw compressor super-charger for the i. c. engine layout of Figure 1; Figure 3 shows a sectional view of the'gearless'screw compressor of Figure 2; Figure 4 shows the screw compressor, and attendant i. c. engine configuration of Figures 1 through 3, with provision for lubricant feed by a controlled bleed of (lubricant) fuel (such as diesel oil) from a fuel injector and/or fuel pump; Figure 5 shows a fuel injector (internal working clearance) bleed, divertable for lubrication of a target component, such as a screw- compressor, of Figures 2 and 3; and Figure 6 shows suitable rotors for use in a screw compressor such as depicted in Figure 2 & 3.

Referring to the drawings, a piston-in-cylinder type i. c. engine 10 has a generally conventional construction of crankshaft, cylinders, pistons, timing gear etc. (not shown).

An intake manifold 12, configured as a plenum chamber, or air- chest, connects with cylinder intake ports-and provides a single, common entry point for combustion air intake.

A super-charger 20 feeds air to the intake manifold 12-and is driven directly from the crankshaft (not shown), or by an independent drive, such as a dedicated electric motor.

The super-charger 20 may in turn be fed by an upstream super- charger 40, in this case, the compressor (stage) of a turbo-super- charger, driven by engine exhaust flow.

An inter-cooler (heat-exchanger) 42 and an after-cooler (heat- exchanger) 44 are disposed between the various compressors and the engine 10, as required. As already noted on page 4, rotor"N"profiles disclosed in UK Patent Application 9610289.2 are particularly suitable for operation without external synchronising means. UK Patent Application 9923522.8 discloses a specific design (& profile) that enhances this still further and in particular by using two male and three female lobes as shown at Figure 6.

The construction of the super-charger 20 is shown in more detail in Figure 2 and comprises a housing 22, with intermeshing rotors 26,28.

An inlet port 23 is located in the lower or underside casing, and an outlet port 24 at the other axial end.

The rotors 26,28 are mounted on respective shafts 27,29, supported by bearings 30, with seals 32 and a drive connection 34.

The construction is"gearless"in that there is no (ancillary) synchronising or timing gear, or other mechanical entrainment or inter-coupling, between the two rotor shafts 27,29-as with a conventional screw compressor.

The rotors 26,28 have complementary helical, marginally-spaced, inter-meshing profiles or forms, in close proximity-or (lightly) contacting, with an intervening lubricant film and/or surface coating (s) (not shown).

The prevailing internal (intervening) air pressure forces upon one rotor-in this case a'passive'female driven rotor 28-are substantially balanced about its axis.

Thus, there is minimal torque upon the passive rotor, arising from, or engendered by, air pressure.

The rotor profile is substantially'involute', at points of (inter-rotor)'contact', or closest proximity.

Moreover, the helix (or twist) angle and the length of rotor provide close intermeshing, over substantially all rotor (relative angular) positions.

This in turn ensures that one rotor, in this case a (active male) driving rotor 26, drives the other (passive female) driven rotor 28-without the need for rotor synchronising drive transmission gear.

Moreover, careful design can eliminate torque reversal of the 'non-driving'or'driven'rotor 28-so preventing'chatter', over substantially the entire operating range.

Omission of timing gear (or other drive entrainment, inter- coupling, relative phasing or synchronising mechanism of conventional screw compressor configurations) saves weight, cost, space and complexity. In a further aspect of the invention, depicted in Figure 3, the rotors and housing are configured for zero (or minimal significant) internal compression.

In this form, the device can strictly no longer appropriately be called a'compressor'.

Rather it could be more appositely referred to as a'screw blower' -or'screw displacer'.

This novel screw compressor derivative, according to one aspect of the invention, exhibits certain similar (fluid-dynamic) characteristics to a Rootes blower-in that it displaces the working fluid, without internal compression.

Such a screw blower would use a housing with a much larger exhaust port 24 than a (conventional) screw compressor-in order to achieve zero (or minimal significant) internal compression-and thus a pressure ratio of unity.

In practice, the angle of'wrap'of the rotor's helixes (ie the helix angle and rotor length) can be chosen so that the blower will function with inlet and exhaust ports which extend (all) around the respective ends of the rciors.

Further, the housing 22, which encloses the'intermeshing'rotors, can simply be extended somewhat beyond the rotor span-in order to provide an end space, that will serve as an inlet port 23, or outlet port 24-that is'end space ports'.

This configuration exhibits advantages in both application and manufacture, including: * Greatly simplified provision of inlet and exhaust ports.

No need for (strict) end-clearance control and attendant manufacturing tolerances.

* No need for (strict) rotor end-play control-but rather allowance for some degree of end'float', in one or both rotors.

* Greater freedom with external connection to such'end space ports'-that is, from any side of the assembly; thereby affording greater freedom in application and packaging.

In this aspect of the invention, any of the various other features disclosed elsewhere herein can still be used-ie gearless rotors, with minimal torque transfer, lubrication, by provision of fluid from injector back-leakage (discussed later) and light alloy, or composite, rotors, with suitable coatings.

In practice, the super-charger housing 22 can be integrated with other engine components, such as the crankcase, cylinder head, water pump housing-in order to reduce, still further, overall engine complexity.

A further variant, depicted in Figure 4, includes a facility for rotor lubrication, in order to reduce wear rates and help sealing -thereby improving super-charger efficiency.

The lubricant may be ingested into the airstream upstream of the super-charger whether or not a turbocharger is present-or may be conducted into the supercharger, via its bearings, thus lubricating them also, or by any other convenient method.

The lubricant helps prevent'dry'inter-rotor contact, that would otherwise inevitably lead t some wear, during acceleration and deceleration, in addition to'normal'operation.

Advantageously, a'very small'-and controlled-quantity of diesel fuel could be diverted, (or bled), from a fuel supply, such from as a fuel injector, or fuel pump, as (super-charger) lubricant.

Such a modest (lubricant ; fuel bleed would effectively be 'consumed'by the engine combustion process-and so would have a negligible effect upon its operation, including its exhaust emissions.

Diesel fuel would be readilv available in those applications where the engine uses this as ils main fuel-so would require no additional maintenance, supply or storage provision.

Diesel is an effective lubricant and has a relatively high boiling point-so would be suitable and effective for such super-charger lubrication duty.

Diesel fuel could be provided by routing the'injector leakage', or (controlled)'bleed', from one or more of the fuel injectors, or indeed the fuel pump itself, of a diesel engine (or gas or gasoline engine, with diesel fuel pilot injection)-thus ensuring an accurately metered source of lubricant, with no additional components or systems.

Generally, fuel injection systems (particularly diesel) exhibit inherent injector'leakage'-because of the design of the internals of the injectors.

Thus, fuel leaks past a narrow clearance between the injector needle and the body surrounding it.

Such clearance is carefully controlled-and a relatively small known amount of fuel is allowed to leak from the injector.

The leakage is usually collected and returned to the fuel tank (or elsewhere in the fuel system) by dedicated return pipes, or fed back to the pump inlet; fuel s not allowed to escape into the environment.

Such lubricant fuel leakage could also be used for some other component lubrication.

One embodiment of this invention uses injector leakage to provide a small, metered supply of a suitab'e lubricant (diesel fuel) to a supercharger.

Figure 5 shows an example of (inherent) fuel injector leakage, which can be used to advantage to provide lubricant fuel for another component-such as a screw displacer, configured as a supercharger, of Figures 2 and 3.

Specifically, an injector pintle 58, has a tapered nose 62, interacting with a tapered seat 63, in a housing body 80.

The pintle outlet 66 communicates with an i. c. engine combustion chamber or precombustion chamber.

The pintle nose 62 is immersed in the chamber 67, which is in turn supplied with fuel, under pressure, by a feed passage 72, from a fuel injector pump (not shown).

The pintle 58 as a whole is biased towards the tapered seat 63, by a spring 76 at its opposite (back) end from the nose 62.

Beyond the nose 62, the body of the pintle 58 is configured as a pintle stem 77, with a generally cylindrical shank, located in a complementary bore 78, in the body of the injector housing 80.

Injection pressure (from the injector pump) lifts the pintle 58 against the spring 76 thereby opening the fluid tight seal 62,63 and allowing fuel to flow though the annular gap 66.

Although a close-toleranced inter-fit is provided between the pintle stem 77 and bore 78, some bypass'leakage'of lubricant fuel is allowed.

Specifically, an internal'back leakage'path 83 is inherent- from the head chamber 67, between the pintle stem 77 and the walls of the bore 78, to the'back'end of the pintle 58.

This (controlled)'leakage'83, of lubricant fuel, is'captured', by an appropriately adapted internal configuration-albeit not detailed, but indicated generally by reference 85, and is returned to the fuel tank or other part of the fuel system.

In this aspect of the invention this inherent leakage is advantageously transferred to the screw blower/displacer to provide a useful lubrication function.

Modest, even'minuscule', quantities of fluid in compressors of a number of types, whether vane, screw, Rootes, or otherwise, can significantly improve compressor performance, by effectively reducing running clearances and hence working fluid leakage.

Some supplementary cooling action may arise-which may be beneficial in lowering working gas temperatures and so raising charge density.

For some industrial compressor applications, fluids are intentionally introduced for various such purposes-and subsequently extracted downstream.

In that case, the improvement in efficiency is sufficient to offset the extra cost and complexity involved in supplying and removing the fluid.

Conversely, clearances may be increased, in order to reduce costs -if fluids can be introduced to effect or promote sealing, without loss in performance.

The back leakage fuel quantity is so small that it will make an insignificant difference to engine performance, and because it is the same as the engine's main fuel, its combustion in the engine will produce the same products and thus cause no significant additional environmental problems.

The lubricant may be fed into the intake of the supercharger, or it may be beneficial to pass it through the rotor bearings, in order to lubricate them also, on its way into the supercharger, or via any other suitable route.

Figure 4 shows an embodiment of this for an i. c. engine as in Figure 1-although the principe is applicable to other configurations of positive displacement i. c. engine, such as rotary types.

Corresponding parts are given the same references and will not be described again.

Again, a primary supercharger 20, configured as a screw compressor according to one aspect of the invention, is fed by an upstream supercharger 40, in this case the compressor of a turbo- supercharger.

The forced induction system may be provided with as many stages of compression as necessary.

A fuel injection system controls the amount of fuel supplied to the engine and comprises a pump 50, injector supply pipes 52, injectors 54 and injector leakage-return lines 56.

The supercharger 20 is provided, via line 56, with a metered supply of diesel fuel, to act as a lubricant.

As described above, some excess diesel fuel is typically produced as waste by the fuel injectors 54-and is referred to herein as 'leakage'.

Various types of fuel injection system use similar principles, but the majority exhibit controlled injector leakage.

The leakage is usually collecte and returned-either to the upstream side of the injector pump, or to an engine fuel tank.

In this aspect of the present invention, at least some of the leakage is instead directed, via feed line 56, to a primary supercharger 20, as described-and/or to lubricate some other engine componentry.

COMPONENT LIST 10 i. c. engine 12 intake manifold 20 super-charger 22 housing 23 inlet port 24 outlet port 26 rotor 27 shaft 28 rotor 29 shaft 30 bearing 32 seal 34 drive connection 40 turbo-supercharger 42 inter-cooler 44 after-cooler 50 fuel pump 52 supply/feed line 54 fuel injector 56 fuel bleed line 58 injector pintle 62 (tapered) nose 63 (tapered) seat 66 pintle outlet/head 67 chamber 72 feed passage 76 spring 77 pintle stem/shank 78 bore 80 injector housing 83 lubricant/fuel leakage (clearance) 85'back leakage'passage