CA1210937A - Pre-compressing apparatus for internal combustion engines - Google Patents

Abstract

ABSTRACT
This invention relates to apparatus that compresses the combustion air, or the air-fuel mixture, prior to being admitted to the combustion chambers of internal combustion engines. The invention is applicable to engines having multiple combustion chambers. The pre-compression is cur-rently practised as an option in piston-type engines by means of supercharging compressors; whereas in dual-shaft gas turbines - being indispensable - it is achieved by the gasifier charging compressors. Both types of compressors are driven by turbines utilizing a partial expansion of the whole volume of burnt gases produced in the engine.
The main feature of the invention is that only a fraction of the total volume of produced burnt gases is employed in the driving turbines. In a given multicylinder engine, said fraction is delivered as active exhaust gases - through a separate exhaust system - by a selected number of cylin-ders from the engine's set of cylinders. The active condition of the gases is achieved by suitable adjustments of the mechanism commanding the operation of the selected cylin-ders. In the remaining cylinders, the combustion products undergo a feasibly complete expansion. Likewise in a dual-shaft gas turbine the fraction of burnt gases delivered to the gasifier turbine as working substance, is obtained from a selected number of combustors, from the engine's combus-tors set. The burnt gases from the remaining combustors are fed to the power turbine to produce the power output.
The implementation of the invention will result in: an eco-nomical use of the thermal energy developed by the combus-tion of the fuel, a better control of the pre-compressing apparatus, and a reduced size of the driving turbines that will handle a diminished volume of working substance.

Description

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G~N~RA~ DISC~S~ION
In the present disclosure 'ipre-compressingl' denotes the operation by which the pressure of the combustion air - or of the air-~uel mixture - is purposely boosted prior to being admitted to the cylinders or to the com-bustion chambers of the inter~al combustio~ e~gines,said operation being carried out by an appliance or a mechani~m that may be considered as a separate component of the engines . ~he supercnargers attached to engines of the piston type - including possibly engines equipped with rota~y pistons - are typical pre-compressing appli-ances; but in the stxict sense of the word, the charging compressors which supply the combustion air to the com-bustors of the ga~ turbines shall also be considered as pre-compressing de~ices.
~his invention relates to sy~tem and apparatus designed to accomplish the pre-compressing operation as defined in the precedi~g paragraph, whereby the pre-compressing apparatu6 can form either a~ optional or an obligatory csmponent of the respecti~e engine6.
In piston-type e~gines supercharging results, primarilyl in an improvement of the engines' volumetric efficiency;
but the combustion of the fuel i8 also improved, and a higher mean effective pressure is achieved. Supercharging is also a convenient means to attain an increa~ed pres-

2~ sure ratio without i~creasing the stroke-to-bore ratio - and the volumetric compression ratio - of the engines.
~hus more wor~ output is obtained in an economical way from an engine of a given ~i~e.
~specially useful i8 the effect of supercharging in com-pression-ignition engines. In order to carry out effec-tively the thermodynamic cycle, these engi~es, when ~.

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operated with suetion of atmospheric air, require ~olu-metric compression ratios of more than 12 : 1. When the combustion air is supplied pre-co~pre~sed, the volumet-ric compression ratio inside the cylinders can be redueed to less than 9 : 1 .
On the other hand, ga~ turbines must be equipped with charging air compressors which, unlike the optional super-ehargers of the piston-type engines9 form a necessary in-tegral component o~ the engines.
Regardless of the fact that, in piston engines the sup-plied supercharged medium is further being compre~ed before combu~tion ta~es place, while in the ga6 turbines the eharged compressed air is used In the combustion cham-bers without additionæl compre~sion, both kinds o~ pressure boo~ters - i.e. superchargers and charging compressors -perform the ~ame function of pre-compressing the medium being supplied to the engines.
PRIOR AR~
~he supsrcharging deYices currently in use co~sist in most cases of rotary compressors driYe~ at hig~ speeds by means o~ gas turbi~es, whereby the engines' burnt gases provide said driving turbi~es with the needed motive power.
~ow-speed blowers are used with some large diesel engines, the blowers being drivan by the engines' shaft through appropriate transmissions.
Pre-compressing devices of the injector-compressor typ~
ha~e found as yet little aeceptance. ~he Canadian Patents 1,112,055 and 1~114,622 is~ued in 1981 to John J. Haiman comprise injector-compressor supercharger~ u~ g as motive power superheated steam, which is generated by re-co~ering heat that otherwise would be wasted.
The internal combustion piston-type engines ~upply the lZ1~31~

burnt gases to the superchargers' t~rbines in the form of exhaust gases~ that is, after said gases have ~er~ed as working substance in the engines' worki~g space. ~he exhaust ga~es must, however, be delivered to the turbines with such ~alues of pressure and temperature as would develop by expansion the required mechanical work. ~his imposes arTesting the adiabatic expansion within the en-gine at a level of higher enthalpy than would be the ca~e in the absence of supercharging. In the following descrip-tion, theæe higher enthalpy gases will be called "activeexhaust gase~", as opposed to the gases expanded to the lowest feasible conditions which will be designated as "spent e~haust gases".
As ~or the gas turbines, it should be noted that there are two categories of such i~ternal combustion engines: the single-shaft and the dual-shaft gas turbi~es. Although the present invention is not applicable to single-sha~t gas turbines, this kind of engine is being examined here in comparison with the dual-shaft turbines.
In the single-sha~t engine, the hot combustion gases gen-erated in the combustors enter the turbine and are sub-jected to feasibly complete expansion, the ~eat be~ng con-verted to me~hanical energy. ~his energy is transferred to the engine's single shaft, which delivers the engine's power output and, at the eame ti e, drives the engine's charging compressor. ~he right amount of mechanical work required to pre-compress the com~ustion air is thus auto-m tically given up by the turbine~s shaft.
~he dual-sha~t turbine is composed of two parts: a gasi-fier section a~d a power section, each section ~aving itsown turbine and its own shaft. ~esideq its turbine, and mounted on the same shaft, the gasifier includes the air ~21~193~

charging compressor~ ~he shafts of the two sections are independent from each other; but the casings of the tur-bines may be joined together, t;he outlet of the gasifier turbine being shaped to form the i~let of the power tur-bine, thus obtaining a two-stage arrar.gement of turbines.
The burnt gases produced in the combustors' system are suppli9d to the first-stage turbine, wherein they deliver the motive power required by the charging compressor. After being used in this ma~ner, the still active gases are fed to the second-stage turbine to continue acting as a worX
ing substance, thereby producing an usable power output.
~he thermal energy developed by the combustors must sat-i~fy the successive demands of the two turbines.
~RIE~ DESCRIP~I0~ 0~ ~HE I~V~I0~
It is noteworthy that in the pre~compressing systems cur-rently in u e the whole volume of burnt gases produced i~
the engines' combustion chambers, or combustors, furnishes the motive power for the deYices which drive the air com pressor~ ~his invention presents the important feature that only a fraction of the total volume of burnt gases is employed as working substance in the turbines used as dri~-ing means for the pre-compressing de~ices.
The invention i9 applicable to engines with multiple ccm-bustion chambers - namely, to multicylinder piston-type en-; 25 gines, and to gas turbines equipped with multiple combustors;
the term "multiple" being interpreted as "more than one".
For piston type engines the fraction of burnt gases employed to attain the pre-compression of the combusticn air - or the air-fuel mixture - will be obtained from selected cyl-inders out of the cylinders composing the engine.
~or dual-shaft gas turbines the needed fraction of burnt gases shall be delivered by a selected number of combustors 9~7 out of the total -number of the engines' set of combustors.
~he ob~ects and the features of the invention sum~
marized hereinbefore are illustrated in the accompany-ing drawings, i~ which:
~igure 1 is the generic flow diagram of a piston en-gine equipped with a supercharger.
~i~ure 2 is the diagram of a single-shaft gas turbine.
Figure 3 is the diagram of an usual dual-shaft turbine.
~igure 4 is a ~-s diagrzm combining two alternative cyclas of an Otto engine.
Figure 5 is a ~-s diagram combining two alternative cycles of a diesel engine.
~igure 6 is a hypothetical ~-s diagram of the cycle of a single-~haft, open style gas turbine~
~igure 7 shows two diagr~ms of the pressure developed inside the selected cylinders with adapted engines' cycles.
~igure 8 3hows an Otto en~ine having ten cylinders in line, one cylinder being selected to supply the motive power to an injector-compressor.
~igure 9 represents a diesel engine havig five cylin-ders in line, two of which are selected to supply the ex-haust ga~es for driving the supercharger.
~igure 9a is a lateral view of fi~-ure 9.
~igure 10 is a partly exploded view of a V6 diesel en-gine, whereby the bank of three cylinders at one side delivers the exhaust gases which drive the air compressor.
~igure 11 shows a 4-cylinder Otto engine with one cyl-inder chosen for supplying the motive power to the super-charger.
-30 ~igure 12 shows a dual-shaft gas turbine having a set of nine combustors, three of whic~ supply the working substance to the gasifier turbine.

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~ igure 12a shows an alternative arrangement of the components of figure 12.
The figures 1 to 3 illustrate -- with the aid of the re-spective ~egend - the evolution of the burnt gases being 5 aocomplished in the three categories of engines mentioned in the Prior Art. ~he figures show how, at some stage of the evolution, the whole volume of burnt gases produced in the combustion chambers, or ~n the combustors systems, furnishes the motive power which dri~es the air compres-sors. Although in some superc~qarging applioations, cur-rently in use, part of the active exhaust gases may occa-sionally be diverted through ~ bypass valve - called a wastegate - before entering the compressor's driving turbine, the turbine is designed to admit all of the burnt gases produced; and all of the engine's exhaust gases are condition~d for being utilized as working sub-stance in the turbine. A prolonged e~cape of part of this working substance through the appropriately named bypass valve means wasting potential thermal ener~y. In figure 1 a gas turbine is shown as the supercharger's driving means;
however, any other device capable of converting thermai energy to mechanical work may be used instead.
As shown in figure 2, the single-shaft gas tur~ine has only one appliance which utilizes the engine's burnt gases. ~he dual-shaft gas turbine of figure 3 h~s a first stage turbine - the gasifier turbine - which drives the air compressor, and a second 3tage turbine - the power turbine - which gen-erat~ the plant's power output.
A~A~YSIS OF ~RE-CC~R~SI~G SYST~S
The following tentative analysis has been set up with a view to evaluate the pre-compressing systems. Use will be made of the T-s diagrams presented in the figures 4 to 6.

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~he diagrams used in theoretical dsmonstrations represent ideal thermodynamic cycles, whereby in successive stageæ -an ideal wor~ing substance: i8 compressed; receives an amoun~ of heat; eonverts part of the heat to mechanical work; and rejects the remainder into t~e surroundings. ~he four cardinal points (a~, (b), (c), (d), of the T-s dia-grams, mark the inceptio~s, respectively the ends, of the successive stages. ~he cycles achieved in actual engine~
deviate i~ many respects from the ones presented here; the deviations will, however, not affect essentially the con-clusions of the analysis.
~he analysis is based on the following as~umptions.
- ~he fuel used by the various types of engi~es is a petro-leum distillate of suitable quality having a calorific power of 9500 kcal/kg . ~he combustion o~ one kg fuel is being considered in the computations~
~ ~he theoretical quantity of air required for the complete combustion o~ 1 kg fuel is about 15 kg ; however, practi-cally the combustion air supplied to any type of intern~l combu~tion engine exceeds the theoretical amount. It iB
assumed that the quantity of air supplied to any pi~ton-type engine i8 17 ~ ~ kg fuel, and consequently 18 kg of burnt gases will result from the combustion. Similarly, for all gas turbines a supply of 20 ~g air for each kg fuel consumed has been assumed, whereby 21 kg burn~ g~ses are generated.
- At point (a) the combustion air has a temperature of 290K (17C) and a pressure o~ 1 ata; an eventual pressure drop due to air filters and other induction devices is disregarded.
- A back pressure of 1.2 ata is assumed to be exerted by the Eilencers or ~ufflers on the exhaust system.

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~Zl(~,37 ~he values of the temperatures and enthalpies attributed to the points (b), (c), (d), and (a')~ (b'), (c'), (d'), resulted from trial-and-error computations; they are to be considered as plausible within the frame of this tentative analysis. WherP used, practical ~alues of efficiencies have been adopted ~rom known reference sources.
The T-s diagram of figure 4 exemplifies a cycle achievable in an Otto engine. ~he diagram i8 actually a combination of two alternative cycles: one ~or the engine without superch~rger, and the other for the engine equipped with a turbocharger. The non-supercharged engine has a ~olumet~
ric compression ratio of 8 ~ he supercharger is aB~
~umed to pre-compress the combustion air to a pressure of 1.4 ata . ~aking ad~antage of the f~lling of the cylinder with pre-compressed air, the supercharged engine is de-signed for the lesser vol~metric compression ratio of 6 : 1.
~he diagram (a-b-c-d) represents the cycle of the con~en-tional, non-supercharged, engine. ~he diagram (a'-b'-~'-d') iB the cycle achieved in the ~uperch~rged engine, and the diagram (a-a' d'-d) is the cycle performed by the turbine of the turbocharger. With the chosen volumetric compres~
ratioæ of the reæpectiYe engines and the pre-compression of the air to 1.4 ata ,the resulting temperatures at the pointæ (b') and (c') differ only slightly from the values at the points (b) and (c), justifying the superposition of the diagrams.
In all internal combuætion e~gines the mechanical work i8 produced by the - nearly isentropic - expansion of the working sub~tanee. ~he difference i~ the enthalpy of the u~it weight of the combustion products i.e. of the work-i~g substance - between the inception and the end of the expansion is a measure of the available thermal energg _ ~. _ ~2~93~

bei~g con~ert,ed to mechanical energy. $he difference of the enthalpy ~alue characteri~ing the expansion performed in a gi~en cycle will be called the ent~alpy span.
In the cycle (a-b-c-d) of figure 4 the enthalpy span be-tween points (c) and (d) would equal 433 167 = 272 kcal per kg working substance. ~he thermal energy available for the cycle (a'-b'-c'-d') of the supercharged engine will be, evidently, less than that; this cycle's enthalpy span being equal to 434 - 202 = 232 kcal/kg working substance.
~he turbocharger is composed of a ga~ turbine and a rotary compressor. ~he thecretical amount of energy required to boost the pressure of 17 kg air from 1 ata to 1.4 ata is 17 x 7 = 119 kcal . ~he ideal efficiency of the cycle (a-a'-d'-d) is 0.23 . ~he combined mechanical efficiency f the turbine-compressor unit can be estimated to be 0.84;
the overall efficiency of the turbocharger would then a~-sume the rather low value of 0.23 x 0.84 - 0~19 . ~he re-sulting energy demand o~ the pre-compressing device iB
119 : 0.19 626 kcal/kg $uel. Related to the expanslon energy that would be avallable in the co~ventio~l engine, amounti~g to lB ~ 272 - 4896 kcal/kg fuel, the turbocharger energy demand represen~s about 1~ ~. Not all of this energg demand i8 to be regarded as a 109s for the ~upercharged engine; the e~uivalent of 119 kcal is reco~ered as a posi-ti~e work i~pu~ in the process.
I~ like ma~ner with that used i~ figure 4 , a ~-s diagram has been traced in the figure 5 by superposing two hypothet-ical diesel cycles. one being the cycle of a conventional non-supercharged engine, and the other representing the alternative cycle of a supercharged engine.
If the re~uired ~olumetric compression ratio of the co~ven-tional engi~e i8 15 : 1 , a~d if the turbocharger o~ the ~ 3~

alternative engine pre compresses the combustion ~ir to a pressure of 2.2 ata, it will be possible to perform the compression-ignition cycle i~ the latter engine with a much lower ~olumetric compression ratio. In fact, adopting for the supercharged engine a volumetric ratio of 805 : 1 instead of 15 : 1 , almost unchanged conditions at poi~t (b) will be obtained for both cycles. Starting from point (a~ with 1 ata and 290~, the pressure-temperature values attained at point (b) in the conve~tional engine through isentropic compression will be 44.3 ata and 857K . In the case of the superchQrged diesel, the starting point of the cycle per~ormed inside the cylinders iæ moved to (a') whereby the combustion air has the pressure of 202 ata and the temperature of 36~K . ~he pressure at point (b) in the supercharged engine equals 2.2 x 8.51-4= 44 ata and the corresponding air temperature is 363 ~ 8.5~4-854E . Since the heat input and the air e~ces~ are the same in both alternatives, the conditio~s of the working substance at point ( G ) are also unchanged.
~he superch~rger of the diesel engine is also composed of a gas turbine a~d an air compressor. ~heoretically 17.8 kcal are needed for raising the pressure of ~ kg air from 1 ata to 2.2 ata; the 17 kg air supplied for the combustion of 1 kg fuel will require 303 kcal . Due to the higher ~em-perature of the active exhaust gases 3 the ideal efficiencyof the driving turbine i~ appreciably higher than that of the similar device of the Otto engine; a value of 0.31 has been determinad. Similarly, an overall efficiency equal to 0.31 x 0.84 - 0.26 is assumed for the diesel turbocharger ~0 and it6 energy demand results as 303 : 0.26 = 1165 kcal.
~he pre-compression of the combustion air in the supercharg-er will demand about 25 ~ of the total e~panæion energy ~ 9;~7 that would be a~ailable in the con~entio~al non-supercharged e~gine, which ~mount~ to 18(420 - 164) - 4680 kcal/kg fuel.
~o be noted, agai~, that ~ot al:L of the 1165 kcal dema~ded represent a lo~s for the supercharged engine.
In all gas turbines the thermodynamic cy¢le is performed wit~out any compression of the combustion medium taking place inside the combustion chambers. ~he pressure of the air being charged to the combustoræ' systems is e~feotively the pressure under which heat is supplied to the cycle.
Although in certain applications pressure ratios aq high as 14 : 1 may be used, ratios in the range 8 s 1 to 10 : 1 are uæual with current charging compressors. In the present ~ ysis the co~se~ative ratio of 8 : 1 h~8 been chosen.
he ~-s diagram of figure 6 illustrates the hypothetical cycle of a single-sha~t gas turbine of the type represented in figure 2 .
As aææumed i~ the premises of the analysis, the combustion of 1 kg fuel shall produce 21 kg burnt gases. Although the cycle represents the composite operation of two mecha~isms9 the turbine and the ch~rgi~g compressor, we can consider separately the func~ions of said two components. ~he t~er-mal e~ergy to be con~erted to mechanical work in the tur-bine is the product of the available enthalpy span by the mass of the working ~ubstance, i.e. 204 x 21 s 4284 kcal per kg ~uel. Assuming that the conversion is accomplished with an efficiency of 0.90 , the equi~alent of 3855 kcal will be disposable at the turbine's shaft. Related to the heat input of 9500 kcal/kg fuel, the efficiency of the tur-bine'~ funotion would be about 0.40 .
An appreciable portion o~ the energy diæpoæable at the tur-bine's shaft i87 however, consumed by the charging compres-sor. Pre-compre~sing 20 kg combustion air to 8 ata require~

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theoretically 20 x 58 = 1160 ~cal/kg fuel. Ass~ing an efficiency of the compression being carried out in the charging compressor e~ual to about 0.90, the compressor's energy demand would amoun~ to about 1290 kcal/kg fuel. Un-like the superchargers used with piston-type engines, the needed motive power of which is supplied by burnt gases being fed to their driving turbines, in the single shaft gas turbines the power demand of the charging compre~sor is co~ered from the mechanical energy already con~erted and transferred to the turbine's shaft which i8 also the compressor's shaft. Since the aYailable thermal energy is being con~erted with an efficiency of 0.40 and the compre~-sion process i8 achieved with an efficiency o~, say, 0.90, the actual efficiency of the pre-compression is about 0.36.
~hi~ value is de~initely higher than the overall efficien-cies attainable with the superchargers o~ the piston-type engines. lhe power demand of the charging compressor equal6 1/3 of the disposable mechanical power at the turbine's single shaft.
~s briefly described before, and illustrated in figure ~ , a dual-shaft turbi~e includes two separate gas turbines, one which drives the charging air compre~sor, and the other which produces the engine's work output. While in the sin-gle-shaft e~gi~e the division between the air compressing work and the work output takes place after all the available thermal energy has been converted to mechanical work, i~
the dual-shaft turbine that division is achieved by con~ert-ing the thermal energy in two euccessive step~, each being per~ormed in the appropriate turbine. A ~-s diagram which would combine the cycles o~ the two separate turbines has not been traced. It ha~, however, been assumed that such a hypothetical diagram - if it were traced - would show _ ~ _ lZ~q~9;~

value of temperature and entha:Lpy at the cardinal points ~a~ (b) (c) (d) of the same magnitude as those determ med for the cycle of the single-sha~t turbine represented in figure 6 . In this figure a poin~ (d') has been marked which is not related to the cycle of the single-shaft en-gine, but which would be ~eaningful for a diagram of a dual~
shaft turbine~ ~he thermal efficiencies of the two turbines depend on the location of (d'3 which would mark the end of the isentropic expansion achieved in the gasifier tuxbine~
and the beginning of the expansion in the power turbine.
~ocati~g the dividing point for an hypothetical dual-shaft plant would be rather speculati~e. ~evertheless, it can be stated that significa~tly lower efficiencies will be at-tained when the potential expansion between (c) and (d) is split in two successive part~ (c)-(d') and (d')-(d) .
~or instance, the dimi~ution of the the~mal efficiency of the first-stage turbine, in the dual-shaft engine, is e~i=
dent considering the shorter e~thalpy span a~d the lesser pressure drop corrgfiponding to the expansion (c)-(d') i~
comparison with the full expansio~ a~hieved in the turbine of the single- shaft engine. In accordance with publishsd reference data, the ~alue of t~e efficiency of the former would not excsed 80 ~ of the value attainable in the lat-ter, bringing the actual efficiency of the ga~ifier'~
pre-compressing apparatus to about 0.29 or less.
Some supercharging system~ comprise an intercooler inserted between the CQmpressOr and the engine's intake manifold.
~he purpose of this accessory is to i~crease the den ity of the combustion air being rammed into the cyli~ders.
Applied to Otto engi~es, the intercooler is suppo~ed to prevent knecking. Since the superchargers produce, in most oases, a lower pre-compression th~n was assumed above, 93~

whereby the temperature increase of the air resulted to be only 29 degrees centigrade, it may be concluded that an intercooler is superfluous. As for compression ignition engines, it was shown above that the increased tempera-ture o~ the pre-compressed air results in a final tem-perature at the TDP quite favorable for the ignition of the fuel, while allowing to reduce substantially the volumetric compression ratio of ~aid engines. In this case, cooling the combustion air is incon~enient; on the contra-ry, some heating of the air might be appropriate to makeeasy the starting of a cold engine. The ~anadian Pate~t No. 917028, issued in 1972 to ~ummins ~ngine Compa~y of U.S.h., proposes thi6 facilitation by burning some fuel within the engine's intake manifold.
~he compression performed in the charging co~pressors of large gas turbines may be divided in two stages, whereby a~ intercooler may be provided between the stages~
Gas turbines - of the single-shaft or dual-shaft types -may also be equipped with regenerators that trans~er heat recovered from the plant's exhaust gaees to the pre-com-pres~ed air.
While these additional devices impro~e, to some degree, the overall efficie~cy of the power plants, it must be noted that they h~ve little, if any, influence on the sfficiency of the turbines which drive the air compressors.
An incon~enience encountered with most superchargers of piston-type engines is their exceedingly high speed. One of the reasons inducing the adoption of high ~peeds is to keep the dimensions of the turbines feaæibly æmall, while handling the large volume of exhaust ga~es flowing through them. Turbochargers spinning at 20000 rpm are successfully adapted to compression ignition engines, but speeds from 3'7 120000 up to 150000 rpm are not unusual for various types of Ctto engines.
As was mentioned before, in some turbocharging systems a wastegate i9 pro~ided in the circuit of the e~haust gases;
there are also systems where ~ second wastegate is inserted in the pre-compressed air duct. ~hese bypass devices are not to be considered a~ regulating devices; but rather as safety val~es, lest their continuous functioning become wasteful.
Summarizing the preceding evaluation, certain drawbacks would appear to be inherent to the precompressing systems currently in use. Most important is the low efficiency of the turbines driving the compressors. A negative effect on the engines' energy balance would result from the relatively high demand of said turbines. An exception to this circum-stance is presented by the single-shaft gas turbine, where all the available thermal energy is converted i~ the plant's only turbine. ~he present invention does not apply to sin-gle-sha~t gas turbines.
Al~A~Y~IS RE~A~ED ~0 ~H~ r~VEX~IC~
As previously mentioned, an important feature of the pres-ent invention is that only a fraction of the total volume of the burnt gases is utilized as wor'~ing substance in the driving means of the air compressing devices.
In order to achieve unchanged pre-compression conditions of the full quantity of combustion air by using only a fraction Qfthe resulting mass of burnt gases, provisions are made to increase the specific work produced per unit mass of said fraction of gases. ~he increase of specific work is attained by extending the adiabatic expansion of the working substance taking place inside the compressors' driving devices.

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~Z1~9~7 ~or piston-type engines the expansion of said fraction of burnt ~ases shall be extended by shifting upwards the point (d') sho~n in figures 4 and 5. ~he increased enthalpy spread and the higher inlel temperature of the working sub-stance will result in a not negli~ible improvement of the thermal efficiency of the respective cycle, thus contribu-ting to perform the needed pre-compression by using less burnt gases.
The efficiency of the production of usable work, from the diminished expansion (c) - (d') in the selected cylin-ders will, of course, oe lessenedO It should, however, be noted that of the two factors which influence the thermal efficiency of a cycle - to wit, the inlet tem~erature of worXing substance, and the enthalpy spread of the performed expansion, it is the former which is the more important.
Consequently more gain is likely to result in the enlarged cycles (a - a' - d' - d) than the resulting loss in the diminished cycles (a' - b - c - d') .
On the other hand, the thermal energy developed by the com bustion of the fuel in the remaining cylinders will be fully utilized, while the efficiency of their cycle - in this case (a - b - c - d) - attains its highest value.
As was explained before, the fraction of burnt gases re-quired to pre-compress the combustion air, in a piston-; 25 type engine, will be obtained from a selected number of cylinders out of the number of cylinders composing the engine.

lZ~(~93'7 ~he active exhaust gases produced in the selected cylin-ders are evacuated through a separate manifold - in the case of a single selected cylinder, through a separate exhauæt pipe - and are conducted to a collecting chamber, from where they are fed as working substance to the tur-bine of the supercharger. After undergoing the expansion i~ said turbine 9 the spent exhaust gases join the equally fully expanded gases that are evacuated from the engine's remaini~g cylinders through their proper exh~ust manifold.
~ince limiting the number of cylinders supplying the work-ing substance to a ~raction of the engine's number of cylinders would result in a lower frequency of the supply - the more so if the engine is of the four-stroke type -the collect~ng chamber is suitably sized and shaped to dampen the irregularities in pressure and flow of said ~upply. ~he collecting chamber per~orms also the function of an afterburner, ensuring the co~plete combustion of the fuel before the active gases enter the turbine.
~he number of cylinders selected to supply the working substance to the pre-compressi~g apparatus depends o~ the type of the engine being eguipped with said apparatus, and on the feasibility of raising to a sufficie~tly high level the enthalpy - i.e. the te~perature and pressure - of the active exhaust gases. It is obvi~us that a larger fraction of the engine's ~umber o~ cylinders, and comparati~ely a higher rise of the e~haust gaseE' e~thalpy are,required to achieve the proposed pre-compression for diesel engines than is the case for Otto engines.
Adequate fractions for the Otto engines would be l/4 or l/3 - also equal to 2/8 , ~/12 .... respecti~ely 2/6 , 4/12 .~.-the numerators being the numbers of selected cylinders, while the denominators stand for the total ~umbers of the _ ~ _ J ~

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en~ines' cylinders. Similarly, practicable fractions for diesels could be 2/5 or l/2 - also equal to 4/lO , 6/15 O--or 2/4 , 3/6 , 4/8 ..,-. Of course, other proportions might be established as suitable depending on the desired pre-compression and on the operating conditions of the engines~
and on the divisibility of the actual number of their cyl-inders. To illustrate the application of the invention, typical engines comprising usual numbers of cylinders will be examined here.
Adopti~g the fraction 1~4 in a 4-cylinder Otto engine, mea~s thRt one cylinder is selected to supply the working substance to the driving turbi~e of the supercharger; with the fraction 1/3 in a 6-cyli~der engine, two cylind~rs are selected to perform together the same function. In the 4-cylinder engine, 4.5 kg of active exhaust gases are thus supplied for each kg of fuel con~umed. I~ the 6Dcylinder engine 6 kg of gases are delivered. In both cases the super-charger shall pre-compress the whole ~uantity of 17 kg air u~ed for the combustion of 1 kg fuel. As shown pre~iously, the equivalent of ll9 kcal i8 required theoretically to boost the pressure of 17 kg combustion air from 1 ata to 1.4 ata . ~he actual energy demand o~ the turbocharger depends on the overall efficiency of the appliance. ~on-ver~ely, this demand must be satisfied by the i~entropic expansion of the working substance supplied by the selec-ted cylind~rs. If in the 4-cylinder engi~e, with the chos~n fraction of l/4 , the temperature of the active gases is raised to 1022 ~ - by shifting the point (d') in fig. 4 -whereby its en~halpy increases to 256 kcal, the energy available for the expansiQn of the gases is 4.5(256-167)5 400 kcal/kg fuel consumed; while in the 6-cyli~der engine with the fraction 1/3 , if the temperature is raised to _ ~z, _ lZ~g93~

967 K - enthalpy becoming 241 kcal - the correspondi~g e~ergy ~ailable is 6~241~167)= 444 kcal/kg fuel.
The conservatively determined overall efficiencies of the turbine-compressor units have the values sf 0~32 for the

4-cylinder engine1 and 0.28 for the 6-cylinder engine; the needed energy being respecti~ely 119:0.32= 372 kcal, and 119:0~28= 425 kcæl . ~oth demands are adequately covered by the available energy amounts.
~wo diesel power plants will be examined, one being a 5-cyl-i~der engine with the chosen fraction of 2/5, and the other a 4-cylinder engine with the fraction 2/4 . ~he quantity of active exhaust ga~e~ being deli~ered to the turbocharger by the selected pairs of cylinders will be 7.2 k ~ kg fuel in the first case, 9 kg/kg fuel in the second case. ~s was calculated before, the theoretical amount of energy needed to compress 17 kg air from 1 ata to ~.2 ata is equal to 303 kcal. Assuming for the 5-cylinder engi~e a temperature of 1115 K - enthalpy 281.2 - the available energy for the cycle of the turbocharger is 7.2(281.2-164)= 844 kcal .
~or the 6-cylinder engi~e with an asæumed temperature of the active exhaust gases of 1070 K - enthalpy 268 - the available energy equ~ls 9(268-164)= 936 kcal.
Determi~ed in the ~ame way as was used for the Otto engines, the turbocharger's overall efficiency o~ the 5-cylinder diesel would have a value of 0.36; similarly, the 6~cylin-der engine having the value of O.33 . ~he turbocharger's dema~d in the first case e~uals 303:0.36~ 842 ~cal per kg fuel consumed, and in the second case 303:0.33= 918 kcal per kg fuel, both demands being satisfied by the abo~e computed amount~ of available energy~
Raising the enthRlpy of the exhaust gase~ being delivered by the eelected cy~i~der~ to sufficien~ly high values is .. . . . . -- . . ~

essential, and to this effect certain adjustment~ have to be made in th~ manner of working of the respectiYe cyli~-ders. To be noted that said adjustments shall apply only to the cylinders selected for supplying the working sub-stance to the pre-compressing apparatus.
If a relatively ~mall raise ~n the enthalpy - me~ning the temperature and the pressure as well - iB needed, main-taining anequally small back-pressure on the engine' 6 exhaust manifold may be sufficient. ~he raises considered in the examples e~ami~ed above require more appropriate adjustments, e pecially with a view to controlli~g the pressure of the acti~e e~h~ust gases.
In fact, in the example of the 6-cylinder Otto engine with the fraction 1/3~ the pressure of said ga~es correlated to the assumed temperature - will amount to 4.4 ata ; wh~
in the 4 cylinder die~el with the fraction ~ 2, the corre-lated pressur~ will be 7 ata . ~he other two e~amples im-plicate even ~igher exh~ust pressures~
~vidently, before the beginning of the evacuation, the pressure of the burnt gases within the selected cy~inderæ
shall not fall below the prerequisite ~alue for performing the turbi~e cycle; ~either shall it exceed unne-~essarily that value. ~he implied adaptation of the cycle taking place in the cylinders iæ te~tatively illustrated ~n the ,25 figure 7 . ~he two diagrams of the figure represe~t the pressure developed in~ide the seleeted cylinders of the engines considered abo~e, in relation with the rotation of the crankshaft measured in angular degree~. Practically, the adaptation will be carried out by modifying the cam-shQfts commanding the intake and exhaust valves and 3 ingeneral, by adju~ting the timing of the admissio~, combus tion, and evacuation of the working substan¢e.
~ , .

l~Zl(~937 As was described previously9 in a dual-shaft gas turbine of current design, the expansion of the bur~t gases generated in the engine's set of combustors is split in two succes-sive e~pansions, performed i~ the two separate turbines of the power plant.
~he innovation of using only a fraction of said gases as working substance for the driving device of the air com-pressor - in this case, the gasifier turbine - is applied to dual-shaft turbines in a specially convenient way.
~he needed fraction of burnt gases shall be delivered by a fraction of the total number of combustors forming the en gine's set; whereby the exp nsion performed in the gasifier turbine is extended from the high enthalpy level attained in the combustors to the practicable lowest level of the spent exhaust gases. ~he remaining combustors will supply the working substance to the pcwer turbine, under ~imilar conditions. As a result of this arrangement~ the thermal efficiencies of the cycles perfo~med in both turbines will be substantially improved in comparison with the relatively low efficiencies of the cycles utilizing the split expan-sion of the burnt gases.
~or the dual-~haft gas turbine there is no need of a col-lecting chamber, the engine's combustors supplying continu-ously the burnt gases in two parallel streams to the sepa-rate distributoræ of the engine's turbi~es~The fraction of combustors selected to supply the working ~ubstance to the turbine driving the pre-compres6ing appa-ratus - i.e. the gasifier turbine - will be determined by the o~erall efficiency of said apparatus. In this ca~e a~
overall ef~ieiency of the same mag~itude as the one that was computed for the pre-compreesi~g process of the single-shaft turbine - namely,equal to 0.36, could very plausibly ~Zl1~93~
be attained. Since the required energy for compressing 20 kg combustion air from 1 ata at the assumed pression ratio of 1 : 8 is, theoretically, equal to 1160 kcal, the energy demand of the pre-compressi~g apparatus will amou~t to 1160 : 0.36 = 3222 kcal/kg fuel consumed. ~his repre sents approximately 1/3 of the thermal energy de~eloped in the complete set of combuætors of the engine by the com-bustion of 1 kg fuel. ~he balance of 9500 - 3222 = 6278 kc~
will be con~erted to usable mechanical work in the power turbine with a comparatively higher efficiency.
~onsequently , 1 out of 3 , 2 out of 6 , 3 out of 9 combustors can be conYeniently selected to supply moti~e power to the gasifier turbine.
In the foregoing description, it has been assumed that the ~uel and the combustion air are equally distrlbuted to all the cylinders~ respecti~ely to all the combu6tors~compri~ed in the engine. Due to the circumst~nce that a selected frac-tion of these multiple combustion chambers supplie6 sepa-rately the working substance to the pre-compres~ing appli-~0 ance, it is possible to regulate the required pre- compreY-sion by separately controlling the thermal energy made a~ailable to said ~ppliance. In this manner, the presæure of the combustion air will be regulated - upwards or down-wards - to æuit various operating requirements of the engme without having to sacrifice unused energy from the burnt gaees produced in those cylinders, or combustors~ that do not supply working ~ubstance to the pre-compressi~g appara-~us .
~urthermore, the aboYe described collecting chamber pro~ided for piston-type engines can be used, i~ need be, to supple-ment the energy supply to the turbocharger by i~jecting ~mall amounts of fuel i~to said chamber, to be readily burnt therein.
It might be advantageous, in certain circumstances, to lower the pressuxe of the acti~e exhaust gases supplied to the turbocharger. ~or example, in the case of a diesel engine with the chosen fraction of 2/5, the pressure of 8.2 ata which would correspond to the prerequisite tempera-ture of 1115 K, migh~ be considered as being too high~
~owering the supply pressure to, say, 7 ata would imply reducing the correlated temperature to 1070 K. But this would result in diminishing the available amount of energy for the turbocharger by 7.2(281.2 - 268) z 95 kcal per k~
of fuel consumed. ~hi~ deficiency can be remedied by having reeourse to the above mentio~ed poæsibility to supplement the energy BUpply. Injecting into the collecting chamber about 0.011 kg fuel for each ~g fuel consumed in the engine will suitably cover the deficit.
In case a relatiYely small pre-csmpression is accepted as purposeful for an Otto engine, the lnvention can be applied in conjunction with an alternstive supercharging de~ice.
~or example, if raisi~g the air pressure in a ratio of, say~
1.15 : 1 meet6 particular operating requireme~ts,an injeo-tor-type compressor can be sub~tituted for the usual turbme dri~en rotary compressor. ~8 represe~ted in the ~chematic figure 8 , an appropriate fraction of bur~t ga~es in the form of active ex~aust gases - i8 supplied to the injector-compressor where ~he a~ailable therm~l energy i~ converted to kinetic energyO After an excha~ge in ma~s momentum with the air being compressed, the mixture of the two media is fed to the engine's cycle. ~he theoretical amo~nt of energy required to pre-compress 17 kg air at the specified ratio of 1.15 : 1 is 53 kcal . Assuming an efficiency of 0.66 for the injector~compressor, the e~ergy demand of the pro¢ess - ~3 -'93~would be 80 kcal/kg fuel consumed. In this e~ample, about 2 kg exhaust ga~es per kg fuel consumed would furnish the needed motive power for the inj~ector-compre~sor. The total mass of burnt gases bei~g the sum of: 17 kg air, 1 kg fuel, and 2 kg recycled burn~ gases - i.e. 20 kg, the fraction of 2 kg exhaust gases would represent 10 % of the total. ~hus 1/10 of the number of cylinders compri~ed in the engine could be selected to supply the active exhaust gases. Prac-tically, 1 cylinder out of ~ , or 10 , or 12 cyli~ders will be selected, it being possible to regulate the energy being supplied, by means of the previously described adjustments.
If 2 kg exhaust gasea are supplied with a temperature of 820 K - enthalpy 201 kGal/kg, p = 2.6 ata - a~d are ex-panded to 1.0 ta - en~halpy 151 kcal/kg, T ~ 624 K - the available energy of 2(201-151) = lOOkcal ade~uately coveræ
the demand of the injector-compressor~ ~he final temperature of the mixture will be slightly higher than tha value eor-responding to the adiabatic compression of the air - name-ly 336 ~, instead of 302 K . Althou~h recycling a fraction of burnt gaæes will result in a dilutisn of the oxygen in the mixed medium, the ratio oxyge ~ fuel available for the combustion will remain unchanged; the chemical reaction will7 however, be activated by the increaæed pressure a~d temperature achieved by the pre-compresæion.
In conclusion to the foregoing descriptions, substantial ad~antages of the invention become appare~t. ~he improved efficiencies of the devices which dri~e the air compressors result in a more economical utilization of the thermal energy. Supplying only a fraction of the burnt gases pro-duced, as working substance to ~aid devices, reduces thasize of the same; lower speeds of ~he turbines can also be used. ~he ~ariant of injector-compressor can be adopted.

t7 DE~ D D~SCRIP~IO~ OF ~EE I~E~IO~
~ he following description refers to the listed fig-ures 8 to 12 of the accompanying drawings which illustrate some embodiments of the invention. It should be noted that, besides the described embodiments, the invention includes implicitly other similar applications where pre-compressing systems and apparatus are designed in accord-ance with the invention's stipulated principles and features.
~he invention includes means de~ised to adequately condition the active e~haust gases, and to control their supply as working substance to the pre-compressing devices.
In the application of the invention to multi-cylin-der engines, the mechanism actuating the intake and the exhaust valves of the cylinders selected to deliver their exhaust gases to the pre-compressing system shall be adjusted by modifying the angular position and the shape of the respective camshaft lobes, in relation to the same elements of the remaining cylinders.
~ prerequisit~e back-pressure on the exhaust of the selected cylinders will be maintained by means of a reg-ulati~g device, installed in the passage of the active exhaust gases.
In the implementation of the in~ention to dual-shaft gas turbines, the quantity and quality of the working sub-stance made available to the gasifier turbine shall be controlled by the fuel supply to the respective combustors.
~he invention is especially adva~ageous for fuel-inaection engines, and for engines using gaseous fuel under pressure, whereby theadmixture of the combustible substance to the combustion air takes place in~ide the combustion chambers or near their intake ports. 3ut the application ~2~ 937 of the invention to carburetor engines is not precluded, in which instances the carburetor may be inserted into the æystem either before, or after the pré-compressing de~ice.
Re~erting to ~igure 8, it can be seen that the selecæd fraction of exhaust gases, cond~tioned as an active medium, is conducted through the ~eparate exhaust pipe 81 to the collecting chamber 82, from where it is fed as working sub-stance to the injector-compressor formi~g the engine's pre-compressing device. ~he injector-compressor consists of two parts: a su~ion chamber 8~ and a Venturi tube 84. ~he 9UC-tion chamber encloses a nozzle which discharges the working substance while converting its available enthalpy to kin~tic energy, whereby a vacuum is also produced that dr~ws the combustion air supplied through the induction pipe 841.
~rom the di~erging section of the Yenturi tube, the air mi~ed with the driving exhaust gases is delivered, with increased pressure, to th~ engine's intake manifold 85.
Said manifold distributes the pre-compressed combustion medium to all ten cylinders of the engine~ The regulati~g valve 843, which i5 actuated through a servomech~nism by the signals sen~ by the pressure sensor 844 7 co~tTols the back-preseure exerted on the exhaust of the ~elected cyl-inder. It is possible to interco~nect the sensor 844 with an electronic control l~n;t 842 which can modify - and even override the signals sent out to the regulating valve 843.
~he exhaust manifold 86 collects and evacuates the e~haust gases from the engine's nine cylinders which do not con-tribute to the working substanoe of the pre-compres3ing device and which, accordingly, utilize the full expansion of their combustion products.
~ he engine of ~igure 8 could be representative af a stationary power plant, UBing a variety of combustible 12~ 37 substa~ces, petroleum distillates, natural gas, or anysuitable synthstic fuel. It may be designed either as a fuel injection engine or as a carburetor e~ui ~ d engine.
In the latter version, in order to avoid pre-ignition of the fuel, the carburetor should be installed in the passage of the pre-compressed air, away from the injector-compres-sor.
In the diesel of ~igures 9 and 9a two adjacent cyl-inders located at the end of the five-cylinder block have been select~d to furnish the motive power to the engine's pre-compressing appliance~ ~he active exhaust gases are conducted by means of the separate manifold 91 to the col-lecting chamber 92, from where they are delivered to the turbine 93 which drives the compressor 94. ~he combustion air is aspirated by way of the induction pipe 941 and is discharged, pre-compressed, through the ram manifold 95 into the ~ive cylinders of the engine. Ihe exhaust mani-fold 96 collects the fully e~panded burnt gases from the three cylinder~ whieh do not contribute to the motive power supplied to the turbine 93. ~his manifold evacuates also the spe~t gases from the turbine 93 that are discharged through the crossover pipe 931. Visible in Figure 9a are also: the regulating valve 98 which controls the back-pres-sure on the exh~ust of the two selected cylinders - being actuated by signals sent out from the sensor 99, and a nozzle 97 used as a mount for a fuel injector. If need be, said injector will supplement the thermal energy made availabls to the turbine 93. A system of baffles, in ~hs collecting chamber 92~ is pro~idsd with the purpose of dampsning th~ prsssurs fluctuations.
In the disssl sngine shown in Figure 10 the fractio~
of 1/2 - actually 3/6 - has besn adopted for the selected lZl~P93'7 cylinders that supply the active exhaust gases. ~he 1/2fraction is suitably materialized in any V-type engine because the cylinders are arranged in two equal banks, set at an angle to each other. In the illustrated example the active exhaust gases are produced by the left bank of cyl-inders, and are led by way of the separate exhaust manifold 101 to the collecting chamber 102, from where they are delivered as working substance to the turbine 103~ ~hi9 turbine drives the compressor 104 which aspirates the com-bustion air through the pipe 100. ~he pre-compressed air is discharged i~to the ram manifold 105 which deli~ers it to all six cyli~ders of the engine. The exhaust gases from the right-hand bank of cylinders are evacuated by their proper exhPu~t manifold 106. ~he spent gases from turhine 103 are conducted to the tail pipe of the ma~i~old 106 by way of the crosso~er pipe 1031. ~he regulating ~alve 108, which is actuated by the signals emitted by the sensor 109~ controls the back pressure exerted on the exhaust system of the left bank of cylinders. At varia~ce with the location of the sImilar valve o~ Figure 9, the valve 108 is mounted at the outlet of the collecting chamber.
~he no~zle 107 is provided for a supplementary fuel injec-tor. A set of baffles 1021, designed ~or quieting the fl~w of the gases, is installed inside the collecting chamber.
~igure 10 shows that the collecting chamber is e~closed in a thermal insulation 1022. Such insulation shall be provided arou~d all the component6 of the pre-compressing systems which contain or carry the active exhaust gases.
In the Otto engine of Figure 11, the cylinder at the left end of the cylinder block has been chose~ to supply the motive power to the turbocharger. ~he active exhaust ga~es discharged by said cylinder through the ~epara~e 12~93 ~

exhaust pipe 111 are conducted to the csllecting chamber 112, from where they are fed to the turbine 113. ~he turb~ne dri~es the compressor 114 which draws the combustion air through the pipe 110 and delivers it, pre-compressed~
to the engine's intake manifold 115. ~he exhaust gases ~rom the remaining three cylinders o~ the engine are collected and evacuated by means of the exhaust manifold 116~ The spent exhaust gases from the turbine 113 are also evacuated through the manifold 116, being conducted to it by the crosso~er-downpipe 119. ~he turbocharger is supported by the centre housing 117 which encloses the sha~t bearing.
Similarly to the preceding examples, a regulating valve 118 actuated by remote control, directed by a sensor 119, will maintain the required back-pressure at the exhaust valve of the selected cylinder.
~he dual-shaft gas turbine æhown in ~igure 12 con-sists of four main compone~ts: an air compressor, a gasi-~ier turbi~e, a power turbine, and a 3et of nine combustors.
The set of combustors iæ composed of two groups: one group, designa~ed by the numeral 121, includes the three combus-tors selected to æupply the worki~g substance to the gasi-fier turbine 123; the other group, designated by 12~, com-prise~ the remaining six co~ustors which deliver their combustion products to the power turbine 126. The gasifier turbine 123 drives the compressor 124 through their common æhaft 125. ~he compressor 124 draws the combu~tion air through the induction passage 120 and pre-compresses it to the required pressure. ~he shaft 127 o~ the turbine 126 can be coupled with a power train, or i~ some other way make usable the power output of the plant. ~xcept for their connecting tubes, the ~ine combustors of the engine are practically iden~ical. ~he combustors' air inlet tubes are ~Z1~937 singly fitted to the ou~let portæ of the ringlike dis-charge end of the compressor. ~e outlet tubes of the combustors are also singly formed to con~ect to the inta~e ports of the distributors of the respective turbines.
Each combustor is equipped with a fuel injector 129, and with a spark plug 130. ~he dual-shaft turbine will be pro-vided with a system of sensors monitoring its operating conditions, the information being fed to a central con-trol unit ~ w~l regulate separately the fuel supply to the two gro~ps of combustors. ~he spent exhaust gases from both turbines of the power plant are evacuated by a common exhaust system 128. It is worthy of note, that locating the gasifier turbine inside the ~age formed by the set of combustors, as shown in Figure 12, results in ~ compast arrangement of the power plant.
The dual-shaft gas turbine shown in Figure 12a is a variant of the engine represented in ~igure 12, whereby the two turbines are arranged parallel to each other, in-stead of being placed co-axially. The mode of operation of this gas turbine is quite similar to the one deæcribed before, and the same numerals designate liXe elements of both variants of the power pla~t.
~ s was mentioned in the foregoin~ presentation of the invention , beside~ the described adaptations of the engines' mechanical parts, the efficient operation of the pre compressing systems re~uires additional controlling means. A full spectrum cf such means and the technique of their application are readily offered by the present state of the art.

Claims (9)

The embodiments of the invention in which an exclusive property or privilege is claimed are defined as follows:

1. Apparatus for pre-compressing the combustion air, or the air-fuel mixture, prior to being admitted to the combustion chambers of internal combustion engines, said apparatus being combined and integrated with engines having multiple combustion chambers, such as piston-type multicylinder engines where each cylinder identifies a combustion chamber, and dual-shaft gas turbines which have multiple combustors, whereby only a fraction of the total mass of burnt gases which result from the combustion taking place within the engines' combustion chambers is utilized as working substance in the driving means of the pre-compressing apparatus, the fraction of burnt gases required to achieve the pre-compression being obtained in a given engine from a suitably selected number of combus-tion chambers - that is to say cylinders or combustors, out of the total number of the engine's combustion cham-bers, and whereby both parts of the total mass of burnt gases - namely, the part produced in the selected combus-tion chambers as well as the part produced in the remain-ing combustion chambers, undergo a feasibly complete adia-batic expansion while being employed to produce mechanical work.

2. Apparatus according to claim 1 for pre-compressing the combustion medium - said medium being defined as the combustion air or the air-fuel mixture being supplied to internal combustion engines - combined with piston-type multicylinder engines, whereby the fraction of burnt gases which is utilized as working substance in the pre-cmpres-sing apparatus is delivered by the selected cylinders in the form of active exhaust gases which result from the combustion products after giving up mechanical work in said cylinders, and which are designated as active because they are conditioned to retain sufficient thermal energy in order to supply the required motive power to the pre-compressing apparatus, the conditioning being achieved by making proper adjustments to the mechanism and to the manner of working of the selected cylinders and by maintaining an adequate back pressure on their exhaust system, said active exhaust gases being conducted through a separate exhaust manifold, or through a sepa-rate exhaust pipe, to a collecting chamber which, for the purpose of dampening the irregularities in pressure and flow of said exhaust gases, is suitably sized and shaped and is equipped with a system of baffles, said collecting chamber serving besides as an afterburner ensuring the complete combustion of the fuel which might be entrained unburnt in the exhaust gases, and being also provided with a fuel injector by means of which, if necessary, fuel can be burnt in order to increase or supplement the thermal energy of the fraction of burnt gases, said frac-tion so conditioned and prepared being fed to the driving means or devices of the pre-compressing apparatus wherein it undergoes a feasibly complete adiabatic expansion, after which it is evacuated and disposed of as spent ex-haust gases.

3. Apparatus according to claim 2, comprising a collecting chamber, a gas turbine which is the driving device of the apparatus, and a compressor which serves as the pressure booster of the combustion medium defined in said claim, said turbine being coupled with said compres-sor to form a turbocharger, the gas turbine as well as the compressor being of the centrifugal type, or of the axial flow type, or of any other rotary type, as will be best suited to the conditions of the fluids involved and to the pressure ratio required to be attained in pre-com-pressing the combustion medium, said pre-compressed medium being delivered to all cylinders of the engine through a ram manifold, while the spent exhaust gases from the tur-bine are evacuated through a crossover pipe which joins the exhaust system of the remaining cylinders - that is, of the cylinders which do not contribute to the working substance being fed to the turbine.

4. Apparatus according to claim 1, combined with a piston-type multicylinder engine, where a relatively low pressure ratio to be achieved in the pre-compression of the combustion medium is accepted as purposeful, whereby the needed motive power for the pre- compression can be supplied by a fraction of burnt gases equal to 1/8 or less of the total mass of burnt gases produced in the engine, meaning that only one out of eight or more cylinders in-cluded in the engine delivers working substance to the pre-compressing apparatus , said apparatus comprising a collecting chamber and an injector-compressor by means of which the combustion medium is pre-compressed while being mixed with the fraction of burnt gases that supplies the needed work in the form of kinetic energy, whereby the relatively small amount of burnt gases is recycled in the thermodynamic process, it being possible that the multi-cylinder engine combined with said apparatus be either of the fuel injection type or of the carburetor type, in the latter case the carburetor being installed in the dis-charge passage of - and away from - the injector-compres-sor, in order to avoid pre-ignition of the fuel.

5. Apparatus for pre-compressing the combustion air, or the air-fuel mixture, according to claim 1, com-bined and integrated with multicylinder internal combus-tion engines, whereby the number of selected cylinders, which supply to the driving means of the apparatus the needed working substance, equals a fraction of the engines' total number of cylinders ranging from about 1/4 to about 1/2 in case the driving means is a gas turbine, and about 1/8 to 1/10 in case the driving means is an injector-com-pressor, the fraction of selected cylinders depending on the type and on the manner of operation of the engines, the working substance being delivered by the selected cylinders as active exhaust gases - that is, as exhaust gases the enthalpy of which is sufficiently high to pro-duce the needed mechanical work by adiabatic expansion in said driving means, whereby the active condition of said gases is achieved by suitably adjusting the angular posi-tion and the shape of the camshaft lobes commanding the intake and exhaust valves of the selected cylinders, and by setting the timing of the evacuation of the combustion products in such a way that their pressure should be not below a prerequisite value for performing the thermody-namic cycle in the driving means, and the prerequisite pressure being maintained by a regulating valve installed in the passage of the working substance to the driving means of the pre-compressing apparatus, said passage inclu-ding a collecting chamber which helps evening the pressure and flow of the supplied working substance.

6. Apparatus for pre-compressing the combustion air according to claim 1, combined and integrated with a dual-shaft gas turbine, where the pre-compressing apparatus forms the engine's gasifier section comprising a gasifier turbine and an air compressor being driven by said tur-bine, the needed working substance for the gasifier turbine being obtained as a fraction of the total mass of burnt gases produced in the engine,said fraction be-ing produced by a selected number of combustors equaling a fraction of about 1/3 of the total number of the engine's combustors, while the burnt gases produced by the remaining combustors is delivered to the power sec-tion of the engine, being utilized in the power turbine, whereby both parts of burnt gases - namely, the part produced by the selected combustors as well as the part produced by the remaining combustors, undergo a feasibly complete adiabatic expansion in the respective turbines while producing useful mechanical work.

7. Apparatus for pre-compressing the combustion air, or the air-fuel mixture, being fed to internal com-bustion engines of the multicylinder (piston) type, prior to being admitted to the engines' combustion chambers, the needed motive power for driving said apparatus being supplied in a given engine of said type by the burnt gases produced in a selected number of the engine's combustion chambers - or cylinders - out of the total number of the engine's combustion chambers - or cylinders - said burnt gases being delivered through a separate exhaust system, which comprises also a collecting chamber, in the form of active exhaust gases, the active condition of which is achieved by proper adjustments of the mechanism command-ing the intake and exhaust valves of the selected cylin-ders and also by an adequate back-pressure exerted on the burnt gases being evacuated from said cylinders, said back-pressure being controlled by means of a regulating valve which is inserted in said exhaust system, and which is actuated through a servomechanism by signals from a pres-sure sensor mounted either on the collecting chamber or on the passage of the pre-compressed combustion medium, said sensor being also interconnected with an electronic control unit that can modify the signals emitted by said sensor, thus regulating the pressure of the delivered burnt gases, upwards or downwards, to suit various opera-ting requirements of the given engine.

8. Apparatus for pre-compressing the combustion medium according to claim 7, whereby, due to the circum-stance that a selected fraction of the engine's total number of cylinders supplies separately the working sub-stance to said apparatus, the pre-compression of the com-bustion medium will be regulated to suit various operating requirements of the engine by controlling separately the fuel supply to said fraction of cylinders , it being also possible to supplement the thermal energy made available as motive power to the pre-compressing apparatus by burn-ing corrective amounts of fuel through a fuel injector mounted on the collecting chamber of the active exhaust gases.

9. Apparatus for pre-compressing the combustion air fed to internal combustion engines of the multiple com-bustor dual-shaft gas turbine type prior to being admitted to the engines' combustors, said apparatus forming the gasifier section of said gas turbines, whereby the needed motive power for driving said apparatus is supplied in a given engine of said type by the burnt gases produced in a selected number of combustors out of the total number of the engine's combustors, while the burnt gases produced in the remaining combustors are delivered to the power section of the engine to be utilized as working substance in the power turbine, the total mass of burnt gases that results from the combustion of the fuel being devided in this way in two separate streams of high enthalpy working substance, each of these streams undergoing a feasibly complete adiabatic expansion in the respective turbine while producing useful mechanical work, said dual-shaft gas turbine being also provided with a system of sensors monitoring its operating conditions and feeding the infor-mation to an electronic control unit that will regulate the fuel supply separately to the two groups of combustors which deliver the two streams of working substance, but in such a manner as to attain a correlated operation of the engine.