Note: Descriptions are shown in the official language in which they were submitted.
~ 9 3,~
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 ~.
~ ,3~
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
-- 2 --
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
-- 4 --
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.
-- 5 --
~,
3~
~ 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.
-- 6 --
9 3~
~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.
', !:
~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~
~ ~ .
lZ~ 93'~
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.
- 15 -
, ~
~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.
-- 16 --
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 ~
9~3'1~
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
.. . . . . -- . . ~
L6937
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
- 25 -
~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
26 -
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
- 27 -
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
- 28 -
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
- 29 -
~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.
- 30 -
