GB2435902A

GB2435902A - Air-cycle refrigerated boosted intercooling of i.c. engines

Info

Publication number: GB2435902A
Application number: GB0700686A
Authority: GB
Inventors: Peter John Bayram
Original assignee: Individual
Current assignee: Individual
Priority date: 2006-03-09
Filing date: 2007-01-15
Publication date: 2007-09-12
Also published as: GB0700686D0

Abstract

Boosted intercooling requires high pressure turbocharging such that with a much higher charge air temperature the intercooler removes much more heat than conventionally. This extra cooled, over-boosted charge air, eg from compressor 2 of a turbocharger via an air-air intercooler 4, is air-cycle refrigerated by pressure reduction via an expander (cooling) turbine/reverse-acting supercharger 7, that outputs power at output shaft 9 when 'braked' via an underdriven belt-and-pulley drive to the engine crank, or a generator plus battery pack. A CVT belt drive enables a positive displacement supercharger to be operated as a throttle such that at less than WOT/underdrive conditions engine vacuum pressure drives the supercharger causing expansion cooling and reducing engine idle rpm. Expansion cooling allows high turbo boost pressures that extract more exhaust energy resulting in expander power output that increases fuel economy and power. High boost charge air temperatures may require face-split radiator/intercoolers, or intercooling with engine coolant via a sub-cooling circuit, or pre-radiator.

Description

-1 -2435902 MINIMISATION OF ENGINE THROTTLE / PUMPING LOSSES plus

AIR-CYCLE REFRIGERATED BOOSTED INTERCOOLING This invention relates to supercharged, or turbocharged and in-tercooled internal combustion engines and the application of air-cycle refrigeration (such as is commonly used with Jet en-gines for air Conditioning of aircraft) to such engines. it also relates to the use of variable speed positive displacement supercharger-s as an engine throttle control device that inputs power to the engine, or a variable speed electric generator.

Conventional intercoolers of turbocharged engines cool the charge air with ambient air directly via an air-to-air inter--cooler, or indirectly via an air-to-liquid inter-cooler connected by a run-around pipe system to a liquid-to-air heat exchanger.

However, this means that the lowest practically achievable charge air temperature in Summer would be in the order of at least 11.0 deg.C (20 deg.F) higher than the prevailing ambient air temperature, even with a very large intercooler, since heat exchange can never be 100% efficient.

Conventional turbocharged engines limit the maximum charge air pressure to prevent pre-ignition which leads to engine damage through excessive cylinder temperatures. However, this limits the amount of energy that can be extracted from the engine ex-haust by the exhaust (cooling) turbine of the turbocharger Adding a pressure reducing expander (cooling) turbine (or re-versed positive displacement supercharger) after an inter-cooler reduces the charge air pressure to the engine inlet manifold when a brake load is imposed on the output shaft of the cooling expander such that the turbocharger can then be operated at a much higher boost pressure than conventionally. Thus, the tem-perature of the charge air entering the inter-cooler will also be much higher such that heat exchange to ambient air, or cool- ing of the charge air, would be much increased. It is this in-creased inter-cooling, together with expansion Cooling pressure drop through the expander that refrigerates' the charge air.

It should be noted that essential requirements for air-cycle refrigeration are: inter-cooler heat exchange, and that expander pressure reduction cooling requires that the expander works'.

Simply blowing a turbocharger into a cooling turbine, without intercooling achieves nothing, except increasing losses.

According to the laws of diminishing returns the point at which increasing the inter-cooler size becomes of limited benefit is increased with this system. This system will also maximise the heat exchange of a physically limited size inter-cooler.

A system characteristic is that as Summer ambient operating temperatures increase, so the system becomes potentially more effective, particularly if it has a Cvi belt drive. As ambient temperature increases the turbocharger boost pressure/charge temperature can be increased (to maintain intercooler heat ex-change) and the expander presst,re drop/coo] ing also increased.

However, this also increases intercooler heat rejection as am-bient temperature also increases, and such effect requires that intercooler and radiator arrangements be carefully considered; suitable Possible arrangements are detailed later.

Similarly, a]imlting'factor is that the cooling/charge air temperature must be higher than would cause manifold freeze-up.

Thus,turbo pressure and/or expander pressure drop would require to be controlled/limited by inlet manifold temperatire, or by laser detection of frost and actuation of a de-frost cycle. In any case, even with a fixed ratio belt drive expander, as turbo pressure increases SO expander pressure drop/cooling increases.

Since the inlet manifold charge air temperattre is much cooler than a conventional system, a significantly higher inlet mani-fold charge air pressure can be used such that even higher turbo boost pressures, and higher cooling turbine' pressure drops, be used. With cooler and higher pressures the density of the inlet manifold charge air is much higher than convention-ally, thus specific bhp/litrp would be significantly increased to allow a smaller and lighter engine (for increased fuel econ-omy), or result in increased total bhp.

At times the cooling expander may cool the charge air below dew-point, but such moisture as S condensed will be discharged from the expander in aerosol form for additional manifold and combustion process cooling, and limited by controls that com-pare dew-point temperature with manifold surface temperature.

The free' power output from the shaft of the cooling expander must be offset by exhaust back pressure losses, but nonetheless heat energy is extracted from the engine exhaust. This power output can be fed into the engine crank via a belt and pulley ttat may also be via a CVT drive. Alternatively, the power may be fed into a battery pack via a belt and pulley to an electric generator (that may operate as a motor if required), and that may be speed controlled to eliminate any requireme for a CVT drive. The CVT drive could be a Van Doorne belt-and-pulley sys-tem, or a Torotrak shaft drive system, or hydraulic system.

The charge air may also be directly watercooled by an air-to-water/liquid intercoojer in marine, or static applications.

Similarly, the charge air may be cooled by a direct-expansion (DX) refrigeration cooling coil(*), or the liquid of an air-to-fluid intercooler refrigerated via a compact DX-to-fjuicj plate type heat exchanger. DX intercooling can be provided from the excess refrigeration capacity that is available from a ve- hicle's A/C (air conditioning) system once the vehicle is way-ing at a reasonable speed and at higher engine r.p.m. (just as turbocharger pressure increases); with possibly a plate type heat exchanger refrigratjon condenser utilising engine coolant (details regarding radiator sub-cooling follow later). if the available A/C refrigeration capacity is insufficient to provide all of the required intercooling duty, the charge air will re-quire pre-cooling by a conventional intercooler. Note: chilling the circulated water of the run-around system of an air-to-water intercooling system is ineffectLlaj because as the water is chilled so the water temperature of the ambient air-to-water heat exchanger is also, reduced, thus reducing heat transfer to ambient air. (*) BEWARE about using a DX refrigeration cooling coil in the engine air intake since any leakage of refrigerant gas will subsequently experience high combustion temperatures.

For example R12 (banned for Impact on the ozone layer) becomes phosgene (mustard gas) in the exhaust: alternative]y refriger-at ion intercool ing should be indirect via chi lied water. There is the Possibility to pump down' liquid refrigerant into a re- ceiver, such that it could be released for a boost in inter-cooling in cases where the cooling expander is overdriven' to operate as a supercharger.

Having a variably speed driven cooling expander enables the ex-pander to be overdrjven' to alternatively act as a compressor supercharger when accelerating from low r.p.m.; reverting to an underdriven' cooling expander as turbocharger pressure kjk-i' at higher r.p.m. In this way the power and torque band width can be both increased and flattened. Such effect can also be similarly obtained by, or be complemented by, the use of hybrid battery power output at low speeds. If required, a downstream intercooler can also be provided.

When power output of the cooling expander to a motor/generator is used to charge a battery pack, and with clutch controlled changeover, once the battery pack is fully charged the output of the cooling expander can then be reverted to the engine.

Similarly, the generator could operate in reverse as a motor Suc.that, .witi clutch controlled changeover, power output from the battery pack can also be transferred to the engine crank.

Having changeover of the power output from the cooling expander would enable the battery pack to be fully charged when cruising at highway speeds, when the engine is operating relatively ef-ficiently, such that the battery pack can then be fully used during urban operation, when the engine would otherwise be op-erating in it's least efficient mode. In this way fuel economy is increased.

With a CVT drive system a reverse-acting supercharger could also be operated as a throttle such that the engine air pumping effect at less than wide-open-throttle conditions would impart a force on the rotating elements of the supercharger, which in turn would be transmitted into the engine crank. Adding a 2-speed clutch, throttle on/throttle off, changeover system would improve throttle response. In this application the supercharger should be a positive displacement type, e.g. a Sprintex' screw -type, Roots-type or a multi-piston swash-plate type, to be effective at idle and low r.p.m. Quick throttle response from idle, and for throttle blipping', can be enabled by having a conventional throttle system that by-passes the cooling tur-bine'. If the system is part of a hybrid power system, battery power could also provide for initial throttle response'. A variation would be to have a light pressure supercharger, with-out a cooling expander and an intercooler, but with a variable speed drive, such that the supercharger would be underdriven at less than WOT (wide open throttle) conditions such that engine air pumping forcefully pulls air through the supercharger, such that power is fed back into the engine crank; higher boost pressures together with an intercooler, can also be used.

A conventional throttle can be used together with a throttle operated CVT driven supercharger to enable additional operating features. For instance, where the famed throttle response of a conventional supercharger may wish to be exploited on occasion, the CVT could be locked into whatever over/underdriye ratio that a driver might wish to select, or even select whilst on-the-fly' (e.g. to suit ambient conditions) and conventional throttle control instigated -in this way the system would be much more readily tunable and could be operated in either a Street' mode or a strip' mode. Similarly, in cold ambient Conditions, upon start-up a conventional throttle could be op-erated as a choke and the supercharger overdriven to thereby add heat to the intake air, also requiring an increase in idle rpm to drive the supercharger to further aid warm-up. By having a supercgarger with an internal bypass with a throttle blade, throttle response from idle can be enhanced by snapping open this bypass on upon wide-open accelerator action.

It should be noted that when using a CVT driven supercharger throttle that at idle, when significant levels of inlet mani-rold vacuum Pl.sure.9ccur, typically in the order of 10 psig, tht significant expansion cooling of the charge air will occur i.e. in the order of 40 to 50 deg F.; rather than the heating effect of a conventional supercharger that can cause engine overheating. To prevent cold weather freeze-up would require the use of an (internal) bypass. The effect of the power output from the supercharger throttle into the engine crank at idle would be such that a significantly lower idling engine speed would be possible, that would also reduce fuel consumption.

Operating the system as a means of throttle control effectively eliminates pumping losses', although there would be efficiency losses through the reverse-acting Positive displacement Super- charger and power transmission system, such that a spark ig-nition engine would still not equal the zero Pumping losses of a diesel; although spark ignition engines have other advantages such as being lighter and having lower exhaust emissions.

Adding adjustably variable impeller guide vanes to a Cooling turbine would increase cooling efficiency and shaft output power when mass airflow to the engine is low. As the angle of the vanes reduce, so also will turbocharger pressure be backed-up to thereby oaximjse the charge air temperature and Similarly intercooler heat exchange (without which the cooling turbine would be ineffectual). Similarly guide vanes can be added to the turbocharger turbine impeller to maximise boost pressure at low engine r.p.m.

Where there may also be one, or two other cooling expanders that are steam/gas powered they may also be mounted on a common output shaft together with the charge air cooling expander, or otherwise connected to the same energy absorbing loads.

With a variable speed cooling turbine' system there is the POSSibility, for cold start conditions, to initially over-drive' the turbine' and pressurige the charge air upstream of a Closed Conventional throttle, such that the engine intake air iS heated (also imposing an increased load on the engine) to assist in the rapid heat up of the engine/catalytic converter.

Since this system is based on having high charge air tempera- tures, such that they may be similar to engine coolant tem- peratures, location of the intercooler must be carefully con-sidered. One solution would be to have the intercooler and engine radiator stacked one above the other, or alongside each other, i.e. face split, or, if row split not completely over-lapping, i.e. the intercooler would extend above (or below) the radiator, I. e.partially face split. Another possible solution would be to have the intercooler cooled by engine coolant via a radiator having a sub-cooling circuit with coolant flow to this circuit in contraflow to the air, or locating the intercooler away from the radiator, e.g. at an Opposite end of the vehicle, or on top of atruck cab.

This invention is flOW described by way of example and with reference to the accompanying Figure 1 semi-schematic drawing.

Figure 1 shows ambient intake air 1 to the compressor turbine 2 of a turbocharger. The high pressure hot charge air 3 from the turbocharger is cooled by the ambient air 5 -to-charge air 3 intercooler 4, and with relatively small pressure drop. The cooled, but still high pressure charge air 6 is pressure re- duced by the cooling turbine 7 such that this pressure redtict-ion and expansion is accompanied by a significant temperattire drop, such that the charge air 8 from the cooling turbine is air-Cycle refrigerated. The cooling turbine 7 is slowed, to cause pressure drop through the cooling turbine 7, by imposing a load on the output shaft 9 of the cooling turbine. The re-quired charge air 8 pressure at the engine intake manifold is obtained by providing an appropriately geared drive connection ratio between the turbine output shaft 9 and the connected power absorbing brake' load. The turbocharger is powered by the pressure of the engine exhaust io from which energy is ex-tracted by the turbocharger cooling turbine 11 resulting in a pressure reduced, expanded and cooled exhaust 12. In the case of the load connected to the output shaft 9 of the cooling tur- bine being the engine crank, it can be seen that at lower en-gine r.p.m. when the charge air 3 pressure is also lower, that also the r.p.m. of the cooling turbine is similarly reduced such that there would still be pressure drop, and cooling, through the cooling turbine.

Figure 2 shows a by-pass around the intercooler, Control valve 13, in sequence with control valve 14 provides for control of manifold air temperature via temperatre sensor 15 to prevent freeze-up and speed Winter starting warm-up. Similarly, sensor can also be used to control manifold air temperature by con-trolling turbo boost pressure and cooling turbine' pressure drop. Pressure sensor 16 inputs can be integrated with temper-ature sensor 15 Inputs to control manifold charge air density.

A laser sensor 17 can alternatively be used to detect frost build up and instigate a de-frost cycle. Other conventional turbocharging control systems including overrun/wastegate val-ves, turbocharger exhaust bypass, etc., may also be used.

As shown in Figure 3, any condensed water in charge air 8 would be centrifuged to the walls of the discharge from the cooling turbine 7 such that it can be separated out by the expansion section 17, and the pressure of charge air 8 used to bleed off this water and spray it onto the intercooler 4. When there is no condensation, only a small amount of charge air 8 would be lost through the bleed line 18 that would, in any case, cool the intercooler 4, alternatively a shut-off valve actuated via a,Iatersfnsor qay, be.,used. Alternatively, such condensate water can be collected and high pressure pumped by a positive displacement pump to be spray injected into individual valve intake tracts, or to cool the engine exhaust/exhaust valves and controlled by (injector) valves for operation at high r.p.m.

A variation is shown in Figure 4 wherein the output shaft 9 of cooling turbine 7 powers a compressor 19 located in the exhaust 12 located downstream turbine 11. The effect of this arrange-ment would be to reduce exhaust 10 back pressure on the engine, such that combustion temperature would be reduced whilst allow- ing a very high turbocharger 11 & 2 pressure that, in turn, in- creases both intercooler 4 cooling and allows increased pres-sure drop through turbine 7 to further increase charge air cooling which would further reduce combustion temperature for the purpose of reducing exhaust emissions and for increased en-gine longevity. Alternatively, this arrangement would enable a higher pressure reduction ratio through turbine 11, and thus a yet higher turbo boost pressure for the same exhaust back pressure, plus increased intercooler 4 and turbine 7 cooling.

Figure 5 shows an arrangement for providing a'sub-cooling?pre_ radiator I8/circuit that also counterfiows coolant 19 to air-flow 20 to increase heat exchange efficiency. This sub-cooled' liquid 21 provides for the lower fluid temperature requirements of an air-to-fluid intercooler, or of a DX refrigeration plate type condenser of a DX intercooling system. The fluid flow 22 into the radiator 23 is high temperature coolant from the en-gine block plus hot return intercooling liquid, or DX condenser fluid, with cooled coolant 24 returning to the engine block. As shown, there is an air break between the two (2) radiators to prevent metallic heat transfer between the radiators; a mono-block arrangement may require the addition of a tube row to the Sub-cooling circuit. Also, as shown, the location of the outlet from the main radiator 23 to the sub-cooling radiator 18 will result in the coolant 19 temperature to the sub-cooling radia-tor 18 being cooler than coolant temperature 24 to the engine, which may requre the addition of tube row(s) to the main radi-ator 23, and save a sub-cooling row. The pipework connnectjon arrangement may be turned upside down where the flow of coolant through the engine block is reversed. it should be noted that a liquid cooled intercooler can be smaller due to the better heat transfer of an air-to-liquid intercooler than an air-to-air in-tercoojer, and can be more flexibly located (there can be more than One>, which together make for easier packaging' in an en- gine bay. The smallest possible intercoolers would be air-to-refrigerated liquid intercoolers, which would also further ease packaging'. Should the excess refrigeration capacity that is available from a vehicle's A/C system (once the vehicle is tra- velling at a reasonable speed) be insufficient, a small conven- tional pre-cooling air-to-liquid intercooler will also be re-quired (which could utilise engine coolant, as could a fluid cooled plate type heat exchanger A/C condenser -which would be small and also be flexibly located for packaging' purposes).

When a positive displacement supercharger is operated via a CUT drive as a means of throttle control, some throttle control inertia is introduced that could be eliminated at idle by by-passing the supercharger, e.g. with an internal bypass, having a throttle' blade that could initially be snapped open upon sharp operation of the accelerator pedal as sensed by a throttle position sensor.

With engines having cylinders in a Vee' arrangement there may often be more than one Cl) turbocharger and one (1) intercool-er, Particularly if the intercoolers are liquid cooled, in any case, it may also be easier to package' multiple small cooling turbines', rather than a single larger unit. Multiple Cooling turbines' could be mounted on one (1), or more common output shafts, and associated drive connections could be derived from both ends of the engine crank. A multiple cooling turbine' arrangement would also assist in evening out the distribution of charge air to the engine Cylinders. In addition, the power output shafts of individual cooling turbines' could be clutch disconnected such that, in effect, there would be primary and secondary turbines' such that when total mass airflow is low the pumping efficiency of a secondary turbine is increased and total frictional losses reduced; it being necessary to shut off airflow to the idling turbine' by a throttle' flap, or poss-ibly braking it's shaft. In the case of an idling turbine' having a separate drive connection, and where there is a var-iable speed drive system, the idling turbine' could be set to come online' at such a rotational speed as to provide Positive boost pressure and the instant throttle response from engine Idling speed that superchargers are renowned for; in such cases, and where the turbines' may be of differing sizes, one turbine' could have a CVI variable speed drive and the idling turbine' a fixed or 2-speed drive, possibly with B mechanical gear drive connection to the engine crank. Simply restoring clutch actuated drive and opening up the 2nd turbine' throttle would, in any case, provide an.*mmediatp throttle' response.

N.B. The word turbine' is used in the context that it is in-clusive of reverse-acting superchargers as well as centrifugal radial-flow turbine expanders.

E XAMPLE CALCULAT IONS

Y' Table Data:

For calculation of air temperatLlre rise, or drop, for a given pressure ratio change. r = pressure ratio, V' = factor.

1.6.142 2.1.234 2.6.311 1.2.053 1.7.162 2.2.250 2.7.325 1.3.077 1.8.181 2.3.266 2.8.338 1.4.100 1.9.199 2.4.281 2.9.352 1.5.121 2.0.217 2.5.296 3.0.365

Example 1 A:

Standard turbo/intercool ing with boost pressure of 100 kPa, ambient 49 deg C (322 deg K), compressor efficiency 65%.

From Y' Table r = 2.0 and Y' 0.217 Ideal, 100%, temperature rise = 322 x.217 = 69.9 deg C Ideal rise / compressor efficiency = 69.9 / .65 Actual rise = 107.5 deg C Ambient temperature + actual rise 49 4-107.5 156.5 deg C Charge air temperature drop through a 70% efficient intercooler = 107.5 x.70 = 75.3 deg C Temperature entering engine = 156.5 -75.3 = 81.2 deg C Pressure entering engine after 7.0 kPa drop thro' inter-cooler = 93.0 kPa

Example 1 8:

Air-cycle refrigerated intercool ing with boost pressure eoo kPa, final charge air pressure 93.0 kPa, cooling expander! reverse-acting supercharg efficiency 60% + data as Example 1.

From V' Table r 3.0 and Y' = .365 Ideal, 100%, temperature rise = 322 x.365 = 117.5 dpg C Ideal rise I compressor eficiency = 117.5 / .65 Actual rise = 180.8 deg C -10 -Pmbient temperature + actual rise = 49 + 180.8 = 229.8 deg C Charge air temperature drop through a deeper intercooler having the same approach factor as the intercooler in Example 2, i.e. the same difference between ambient temperature and intercooler charge air leaving temperature = 229.9 -81.2 148.7 deg C Temperature entering expander = 81.2 deg C (354.2 deg K) From VI Table r 1.93 and V' = 0.204 Ideal, 100%, temperature drop = 354.2 x.204 = 72.3 deg C Ideal drop x expander efficiency 72.3 x.60 Pctual temperature drop = 43.4 deg C Temperature entering engine ai.e -37.8 deg C i.e. charge air temperature is 11.2 deg C BELOW ambient.

The increase in charge air density will give an approx 6.7% power increase compared with conventional intercooling.

For a 225 kW engine the increase = 15 kW Cooling turbine shaft power output @ 100 kPa pressure drop and airflow of 235 1/s = 0.00101 x 235 x 100 = 23.75 kW With power transmission loss of 10% = 23.75 x.9 = 21.4 kW 1'S tii Total air-cycle system power increase = 15 + 21.4 = 36.4 kW From this the increased exhaust system back pressure power losses must be deducted.

If the turbo boost pressure is increased to 250.0 kPa and the cooling expander pressure drop increased to 150.0 kPa, and the same deeper intercooler retained, the following performance data would apply: Actual turbo temperature rise = 190 deg C Air temperature entering intercooler 239 deg C -11 -Pir temperature leaving intercooler = 86.4 deg C Cooling expander temperature drop 43.8 deg C ir temperature entering engine 42.6 deg C Engine power increase from in-crease in charge air density = 20.0 kW Cooling expander power output = 34.2 kU lotal air-cycle power increase = 54.2 kW Less deduction for exhaust losses.

Example 2 P:

Standard turbo/Intercooling with boost pressure of 100 kPa, ambient 18 deg C (291 deg K), compressor efficiency 65%.

From y' Table r = 2.0 and Y' = 0.217 Ideal, 100%, temperature rise = 291 x.217 = 63.1 deg C Ideal rise / compressor efficiency = 63.1 / .65 Pctual rise = 97.1 deg C Ambient temperature + actual rise = 18 4-97. 1 = 115.1 deg C Charge air temperature drop through a 70% efficient iritercooler = 115.1 x.70 80.6 deg C Temperature entering engine iis.i -80.6 = 34.5 deg C Pressure entering engine after 7.0 kPa drop thro' intercooler = 93.0 kPa

Example 2 B:

With air-cycle refrigerated ifltercooling with a boost pressure of 200 kPa, final charge air pressure 93 kPa, cooling expander efficiency 60% + data as Example 2 A. From y, Table r 3.0 and.365 Ideal, 100%, temperature rise = 291 x.365 = 106.2 deg C -12 -Ideal rise / compressor eficiency = 106.2 / .65 Pctual rise = 163.4 deg C Ambient temperature + actual rise = 163.4 + 18 = 181.4 deg C Charge air temperature drop through a deeper intercooler having the same approach factor as the intercooler in Example 2.

<p≥ 181.4 -47.6 = 133.8 deg C Temp. entering cooling expander = 47.6 deg C (320.6 deg K) From Y' Table r = 1.93 and V' = 0.204 Ideal, 100%, temperature drop = 320.6 x.204 = 65.4 deg C Ideal drop x expander efficiency = 65.4 x.60 Actual temperature drop = 39.2 deg C Temperature entering engine = 47.6 -39.2 = 7.4 deg C notional Actual temperature entering engine due to condensation effect = 11.1 deg C dry bulb & 10.0 deg C wet bulb i.e. charge air temperature is 8.3 deg C BELOW ambient.

The increase in charge air density will give an approx 4.7% power increase compared with conventional intercooling.

For a 225 kW engine the increase = 10.6 kU Cooling turbine shaft power output 0 100 kPa pressure drop and airflow of 235 1/s = 0.00101 x 235 x 100 = 23.75 kW With power transmission loss of 10% = 23.75 x 0.9 21.4 kW Total air-cycle system power increase = 10.6 + 21.4 = 32 kW

Example 3 A:

Throttle acting CVI driven supercharger at idle, vacuum press-ure 61 kPa, ambient 55 deg C (328 deg}), supercharger efficiency 60% From V' Table r = 1.6 and Y' = 0.142 -13 -Ideal, 100%, temperature DROP = 328 x. 142 = 46.6 deg C Ideal DROP x supercharger efficiency= 46.6 x 0.60 Actual DROP = 279 deg C Ambient temperature -actual DROP = 55.0 -27.9 Air temperature entering manifold 27.1 deg C

Example 3 B:

Throttle acting CUT driven supercharger at idle, vacuum press-ure 61 kPa, ambient 30 deg C (303 deg K), supercharger efficiency 60% From Y' Table r = 1.6 and Y' = 0.142 Ideal, 100%, temperature DROP = 303 x.142 = 43.0 deg C Ideal DROP x supercharger efficiency 43.0 x 0.60 Actual DROP = 25.8 deg C Ambient temperature -actual DROP 30.0 -25.8 Air temperature entering manifold = 4.2 deg C

Claims

-14 -

CLAIMS

1) A reversible adiabatic expansion engine in the combustion air intake of an internal combustion engine that, when it's output shaft is braked by a speed reducing connection to a power gen-erator, reduces the pressure of the airflow passing through it.

2) A reversible adiabatic expansion engine, according to Claim 1, in which the engine's output shaft is connected via a variable ratio connection to the power generator that, thereby, can vary the pressure change of the airflow passing through it.

3) P reversible adiabatic expansion engine, according to Claim 1, in which the speed of the power generator can be varied and that, thereby, can vary the pressure change of the airflow passing through it.