U.S. patent number 3,902,472 [Application Number 05/362,837] was granted by the patent office on 1975-09-02 for diesel engines.
This patent grant is currently assigned to Societe Anonyme de Vehicules Industriels et d'Equipements Mecaniques. Invention is credited to Yves Baugelin.
United States Patent |
3,902,472 |
Baugelin |
September 2, 1975 |
Diesel engines
Abstract
A Diesel engine comprising a set of starting cylinders and a set
of power cylinders, means associated with the starting cylinders,
which means make the starting cylinders more suitable for starting
the engine than the power cylinders, and means associated with the
power cylinders which means makes the power cylinders more suitable
for running the engine under full power than the starting
cylinders.
Inventors: |
Baugelin; Yves (Louveciennes,
FR) |
Assignee: |
Societe Anonyme de Vehicules
Industriels et d'Equipements Mecaniques (92 Suresnes,
FR)
|
Family
ID: |
26217098 |
Appl.
No.: |
05/362,837 |
Filed: |
May 22, 1973 |
Foreign Application Priority Data
|
|
|
|
|
May 24, 1972 [FR] |
|
|
72.18556 |
Nov 15, 1972 [FR] |
|
|
72.40568 |
|
Current U.S.
Class: |
123/179.19;
123/179.16; 123/198F |
Current CPC
Class: |
F02N
19/001 (20130101); F02M 59/42 (20130101); F02D
41/0082 (20130101); F02D 41/062 (20130101); F02B
2075/1824 (20130101); Y02T 10/12 (20130101); F02B
3/06 (20130101); F02B 2275/14 (20130101) |
Current International
Class: |
F02M
59/42 (20060101); F02N 17/00 (20060101); F02M
59/00 (20060101); F02B 3/06 (20060101); F02B
75/00 (20060101); F02B 75/18 (20060101); F02B
3/00 (20060101); F02h 017/08 (); F02m 059/42 () |
Field of
Search: |
;123/179E,179 L:DIG./ 8/
;123/179R,139ST,179G,198F |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Myhre; Charles J.
Assistant Examiner: Rutledge, Jr.; W.
Attorney, Agent or Firm: Eslinger; Lewis H. Sinderbrand;
Alvin
Claims
I claim:
1. A diesel engine comprising, a plurality of cylinders arrayed in
two predetermined sets consisting of a set of starting cylinders
and a set of power cylinders, and means for simultaneously
supplying fuel to said two sets of cylinders at differing
predetermined flow rates selected to make the set of starting
cylinders operative for starting the engine, at a higher volumetric
compression ratio than the set of power cylinders, while said set
of power cylinders are operative for running the engine under full
power at a lower volumetric compression ratio than the set of
starting cylinders and with maximum efficiency; said means for
supplying fuel to said cylinders including an in-line injection
pump and separate injection means associated with each of said
cylinders, the injection means for supplying fuel to the starting
cylinder supplying fuel thereto at a rate which is less than that
at which fuel is simultaneously supplied to the power cylinders
from the injection means associated with the power cylinders; said
injection means comprising individual simultaneously operated
plungers in the pump for supplying fuel to the respective
cylinders; said plungers each having fuel control profiles of
substantially the same predetermined configuration and being set in
predetermined relative positions in said pump for simultaneous
movement, the profile setting of the plungers associated with the
starting cylinders being different from that of the plungers
associated with the power cylinders.
2. An engine according to claim 1 wherein the number of starting
cylinders is small compared with the number of power cylinders.
3. An engine according to claim 1, wherein the starting cylinders
have an initial or static lead value smaller by a value .alpha.
than that of the power cylinders, the phase displacement of the
injection lead of the starting cylinders as compared to the power
cylinders is established by said in-line injection pump; said pump
having cam means of predetermined contour respectively associated
with said plungers for reciprocating the plungers, said cam
contours being identical and the cams of the plungers feeding the
starting cylinders being dephased by the value .alpha., in order to
reduce the injection lead as compared to the cams of the power
pistons.
4. A diesel engine comprising a plurality of cylinders arrayed in
two predetermined sets consisting of a set of starting cylinders
and a set of power cylinders, and means for simultaneously
supplying fuel to said two sets of cylinders at differing
predetermined flow rates selected to make the set of starting
cylinders operative for starting the engine, at a higher volumetric
compression ratio than the set of power cylinders, while said set
of power cylinders are operative for running the engine under full
power at a lower volumetric compression ratio than the set of
starting cylinders and with maximum efficiency; said means for
supplying fuel to the starting cylinders including separate
injection means associated with each of said cylinders, the
injection means for supplying fuel to the starting cylinder
supplying fuel thereto at a rate which is less than that at which
fuel is simultaneously supplied to the power cylinders from the
injection means associated with the power cylinders; and wherein at
the injection outlets leading to the starting cylinders are
installed non-return valves having a low rate of flow and at the
injection outlets leading to the power cylinders are installed
non-return valves having a high rate of flow.
5. A diesel engine comprising a plurality of cylinders arrayed in
two predetermined sets consisting of a set of starting cylinders
and a set of power cylinders, and means for simultaneously
supplying fuel to said two sets of cylinders at differing
predetermined flow rates selected to make the set of starting
cylinders operative for starting the engine, at a higher volumetric
compression ratio than the set of power cylinders, while said set
of power cylinders are operative for running the engine at full
power at a lower volumetric compression ratio than the set of
starting cylinders and with maximum efficiency; and wherein said
fuel supply means comprises a rotary pump having distribution
passages, for feeding the starting cylinders, staggered in the
direction of rotation by an angle .alpha. with respect to the
setting of passages for feeding the power cylinders whereby the
static injection lead value of the starting cylinders is smaller by
a value .alpha. than that of the power cylinders.
6. A supercharged diesel engine comprising a plurality of cylinders
arrayed in two predetermined sets consisting of a set of starting
cylinders and a set of power cylinders, and means for
simultaneously supplying fuel to said two sets of cylinders at
differing predetermined flow rates selected to make the set of
starting cylinders operative for starting the engine, at a higher
volumetric compression ratio than the set of power cylinders, while
said set of power cylinders are operative for running the engine
under full power at a lower volumetric compression ratio than the
set of starting cylinders and with maximum efficiency; wherein
constrictors having calibrated orifices are situated at inlet ports
of the starting cylinders in such manner that the supply of air to
the starting cylinders is reduced at high engine speeds.
Description
The present invention relates to multicylinder Diesel engines.
It is known that, in a Diesel engine, the volumetric compression
ratio should be high to ensure the initial firing actions when
starting the engine from cold. In normal operation of the engine,
when the coolant has reached its stabilized temperature level
T.degree., a lower compression ratio is favourable from the point
of view of efficiency of operation and durability of the engine. In
point of fact, a reduction of the compression ratio results in a
reduction of the maximum pressure of the cycle, which leads to a
reduction of the mechanical stresses and of the frictional forces,
particularly if the engine is supercharged.
By virtue of this fact, Diesel engines commonly have an
intermediate compression ratio which provides a compromise between
ease of engine starting and the reduction of the maximum cycle
pressure within the cylinder or cylinders.
To eliminate this unwieldy compromise moreover, engines have been
suggested, wherein the or each cylinder offers a variable
compression produced by variation of the unswept volume.
Translatory displacement of the precombustion chamber roof or a
variable distance between the piston gudgeon pin and the piston
crown, render it possible to establish a high compression ratio,
for example about 20, for starting purposes, and to cause this
ratio to decrease in gradual manner to a minimum value slightly
exceeding 10. With such engines which offer the advantage of
flexibility of adjustment, the latter remains chancy on occasion,
and the cost price of the device is not negligible.
The same description, as a compromise, is applicable to the
selection of the time-lag upon closure of the inlet valve.
It is known that the inlet valve is closed several tens of degrees
after bottom dead centre (hereinafter referred to as BDC) during
the compression stroke of the piston. This renders it possible to
assure optimum supply of air at high engine speed at which, taking
the inertia of the flow of induction air into account, a delay or
overlap increases the charge despite the return of the piston. At
low engine speed however, the dynamic action of the air is weak, so
that the inlet closure delay (hereinafter referred to as ICD) is
useless or even harmful. In point of fact air is returned into the
inlet and the true volumetric compression ratio of the cylinder is
reduced by a value linked with the loss in the compression stroke
between BDC and ICD, which correlatively impairs the ease of
starting linked to the intensity of the compression.
A relatively small ICD is adopted, to assist starting, which is
sufficient however for charging at maximum speed. The improvement
in starting, requiring an ICD equal to nought, operates against the
high power output linked to air induction at high engine speed. It
was for this purpose that variable distribution mechanisms have
been devised, in particular being dynamic devices wherein the ICD
increases in continuous manner as the engine speed rises. In this
case too, their cost and ticklish adjustment proved to be an
obstacle to the development of these mechanisms.
In accordance with the invention there is provided a Diesel engine
comprising a set of starting cylinders and a set of power
cylinders, means associated with the starting cylinders, which
means make the starting cylinders more suitable for starting the
engine than the power cylinders, and means associated with the
power cylinders which means make the power cylinders more suitable
for running the engine under full power than the starting
cylinders.
A Diesel engine according to the invention tends to start easily
and has a high performance under reduced mechanical stress
conditions, particularly if the engine is a supercharged
engine.
The starting cylinders may have a higher volumetric compression
ratio than the normal compromise assuring normal power in order to
provide the means referred to above.
The volumetric compression ratio of the power cylinders may be
reduced to a value establishing optimum efficiency. By way of
example, for an unitary swept volume of one litre or a bore value
taken as approximately 100, the compression ratio of the starting
cylinders may exceed 20, the compression ratio of the power
cylinders being approximately 13.
A differentiation may be established between the starting cylinders
and the power cylinders by a different value of ICD. A smaller ICD
than the compromise value assuring optimum charging is adopted for
the starting cylinders, to increase their compression. The optimum
ICD charging value is adopted for the power cylinders, without any
concession to ease of starting.
The differentiation between the starting and power cylinders may be
established by means of their injection system. To this end, the
cylinders of both categories are equipped with an injection system
differing in structure or adjustment. The active injection elements
of the two kinds are preferably comprised and actuated within one
and the same pump unit, and are preferably in-line. The
differentiation between the starting and power cylinders may be
established at the point of the injection pump cylinders.
The maximum delivery injected per cycle into the said starting
cylinders may be smaller than the maximum feed injected into the
power cylinders. This difference may be produced on an in-line
injection pump by several means such as variation of the diameter
of the pistons, of the cam contours, or of the configuration of the
injection headers. To simplify matters, it is preferable to employ
identical in-line pump elements with appropriate setting of the
pump plungers feeding, respectively, the injectors of the starting
and power cylinders with respect to a common rack.
It is also possible to establish a differentiation between the
feeds injected into the two kinds of cylinders by choosing
different return valves, different injection pipe lengths and
diameters, and different calibrations and diameters of the
injection holes, which represent parameters affecting the quantity
and pressure of the flow injected. This is especially useful in the
case in which a rotary pump of the distributor type is employed
instead of the conventional pump comprising in-line pistons.
A differentiation between the starting cylinders and the power
cylinders may be established by the value of the initial or static
injection lead, in order to optimize the operation towards the
starting range or towards the range of maximum efficiency and
power. It is known that, for the starting cylinders, a reduction of
the injection lead improves the initial firing actions and reduces
maximum pressure in normal operation.
This differentiation of the injection lead may be produced at the
point of the injection piston by displacement of the top edge with
respect to the feed opening, by adopting a piston level, an
automatic advance injection header or a cam contour, which are
different in the case of in-line pumps. To simplify matters, it is
preferable to adopt in-line pump elements having identical cam
contours, but whereof the setting is different with respect to
uniform theoretical distribution, meaning that the cams of the
pistons feeding the starting cylinders are dephased to reduce the
injection lead with respect to the cams of the power pistons.
In the case of a rotary pump, it is possible to ring the changes by
irregular distribution of the distributor passages.
Finally, preferably a small fraction of (d) starting cylinders is
selected among the n cylinders of an engine; d and n being such,
for example, that d amounts to at most 1/3rd or better yet to 1/4,
this last proportion corresponding to the most satisfactory form of
embodiment contemplated.
The d starting cylinders will be selected as a function of the
evenness of the firing order of the set, of their particular
location from the mechanical point of view (torsional vibration,
more rigid main bearing at the centre or at the extremities), of
their position with respect to the inlet or exhaust ducts, and
depending on whether they are combined under one and the same
cylinder head or not, in order to concentrate the auxiliary
starting means or to establish the corresponding compression
changes on a common cylinder head.
The flow of air intended for the starting cylinders may be
restricted by shutters calibrating their inlet ports. This results
in a greater flow of air towards the power cylinders, into which a
greater quantity of fuel is injected.
This arrangement renders it possible to secure a gain in power at
full speed without affecting the starting conditions, owing to the
fact that the shutters or baffles which do not impede the passage
of air at the low rates of starting flow at which they produce a
small pressure drop, produce an increasingly substantial pressure
drop at "full-bore" speeds, at which the flow is then effectively
reduced.
Some engines embodying the invention are described below by way of
example, with reference to the accompanying drawings, wherein:
FIG. 1 is a diagrammatic view of a 6-cylinder in-line Diesel
engine, the Figure showing the cylinders in plan and the injection
pump of the engine;
FIGS. 2A and 2B are alternative diagrammatic longitudinal vertical
sectional views through the engine shown in FIG. 1;
FIG. 3 is a view of the injection pump and the non-return valve
associated with one of the starting cylinders, of the engine shown
in FIG. 1, the non-return valve being shown in section;
FIG. 3A is a sectional view of the non-return valve of one of the
power cylinders of the engine shown in FIG. 1;
FIG. 4 is a sectional view of the settings of the injection heads
of the injection pumps shown in FIG. 1;
FIGS. 5A to 5C are developed views of injection heads of injection
pumps;
FIG. 6 is a diagrammatic plan view of the cylinders of a V-8
cylinder engine;
FIG. 7 is a diagrammatic plan view of the cylinders of a
12-cylinder engine having two opposed lines of 6 cylinders;
FIG. 8 is a longitudinal cross-sectional view of a rotary injection
pump;
FIG. 9 is a plan view of the rotary pump shown in FIG. 8 showing
the setting of the injection; and
FIG. 10 is a diagrammatic plan view of an engine and its inlet
manifold.
FIG. 1 shows a Diesel engine having six cylinders 1 to 6, which may
or may not be supercharged. The firing order of the cylinders is 1,
5, 3, 6, 2, 4. The engine may be of the precombustion chamber type,
as shown in FIG. 2B. As shown in FIG. 2B, each cylinder
accommodates a flat-topped piston 7 and communicates with an
injection chamber 8. Alternatively as shown in FIG. 2B, the engine
may be of the direct injection type. As shown in FIG. 2A each
cylinder accommodates a piston 9 having a chamber 10 in its
crown.
Each cylinder is associated with a fuel injector 11P or 11D which
is connected to a barrel 12P or 12D of an injection pump with
in-line barrels. Within each barrel 12P or 12D is accommodated a
plunger 13P or 13D (see FIGS. 1 and 3) whereof reciprocating
displacement is performed by a rotary cam 14P or 14D and whereof
pivoting displacement is performed by a rack 15 which engages with
a toothed portion of each plunger. The rack 15 can be
longitudinally reciprocated as shown by the double-headed arrow A.
Longitudinal movement of the rack 15 causes all the plungers 13P or
13D to pivot.
The cylinders 3 and 4 are for starting the engine under high
compression, whether they are under the same cylinder head, if the
engine comprises three cylinder heads, or if it would be
advantageous for the two starting cylinders having a high
compression ratio to act on the central main bearing, whereas the
cylinders 1, 2, 5, 6 are for operation of the engine at normal
engine power.
As shown in FIGS. 2B and 2A, the differentiation between the
compression and starting ratios is established by a variation of
the unswept volume 16 and of the topdead centre either by reducing
the volume of the chamber 10.sub.4 in the piston 9 in the case of
direct injection or by reducing the volume of the chamber 8.sub.3
incorporated in the cylinder head or by combination of these two
means.
The cylinders 1, 2, 5, 6 have a greater unswept volume 16 and thus
have a lesser volumetric compression ratio. The true compression
ratio of the cylinders 3 and 4 is increased further by adopting a
smaller inlet closure delay of the order of 16.degree. of crank
angle, whereas the power cylinders 1, 2, 5, 6 have an inlet closure
delay of the order of 45.degree. of crank angle.
As for induction, the delivery of fuel in mms.sup.3 /cycle to the
power cylinders 1, 2, 5, 6 exceeds the delivery injected into the
starting cylinders 3 and 4. This result is obtained by separate or
simultaneous adjustment of the facilities inherent in the injection
device.
To obtain these differences in fuel delivery, in particular into
the power cylinder 6 and into the starting cylinder 4 (FIG. 1), the
heads of the plungers 13P and 13D developed views of which are
shown in FIG. 4, are identical but are differently set, the rack 15
engages them at the same time in the same manner.
On the developed views of the pistons 13P and 13D have been plotted
areas 17P, 18P, 19P and 17D, 18D, 19D, which correspond,
respectively in operation at idling speeds, at maximum rate of flow
and under overload, the extremity 20P or 20D correspond to cut-off
of the feed.
The setting of the extremity 20P of the piston 13P with respect to
the fuel inlet orifice 21P corresponds to the value .theta.P for
the maximum rate of delivery 18P, whereas the setting of the
extremity 20D of the piston 13D with respect to the inlet orifice
21D corresponds to the value .theta.D for the maximum rate of
delivery 18D.
The settings of the pistons 13P and 13D thus differ by a quantity:
.theta.P - .theta.D = .tau.
which corresponds to a difference in the flow of fuel by the action
of the profile or helix 22 (FIGS. 3 and 4) whose principle is well
known.
Because of the different settings of the injection elements of the
plungers 13P and 13D. The useful injection stroke for the maximum
delivery rate of the areas 18P, 18D corresponding to the rate of
delivery Q for the power cylinder 6 and to the rate of delivery q
for the starting cylinder 4, has a profile length differing by the
quantity .DELTA..
This simple displacement renders it possible to retain the rate of
overload delivery corresponding to the area 19P, 19D and to cause
the starting plungers 13D to stop at 20D.
FIG. 3 shows a non-return valve 23D for controlling the delivery of
the fuel to the injector 11D of the starting cylinder 4. FIG. 3a
shows a valve 23P for controlling the delivery of the fuel to the
injector 11P of the power cylinder 6.
A modification of the distances 24D and 24P between the passages 40
and the valve seats may represent a complementary means of
establishing a differentiation between the rates of delivery Q and
q. The non-return valve 23D, by virtue of the distance 24D
corresponds to a considerable rate of flow -Q' subtracted from the
quantity delivered by the pump and to a small quantity q injected,
whereas by virtue of the smaller distance 24P, the valve 23P
corresponds to a small rate of flow -q' subtracted from the
quantity delivered by the pump and to a large quantity Q
injected.
This expedient is intended more particularly for application of a
rotary pump and the action may be intensified by a greater
calibration of the injectors 11D installed on the starting
cylinders 3 and 4.
As for injection lead, this may be affected by the cams 14P and 14D
(FIG. 1) and by the features of the profile at the beginning of the
injection (FIGS. 5a to 5c).
It is known that the cams of an in-line injection pump for a
6-cylinder engine are dephased consecutively by an angle of
60.degree. in accordance with the sequence 1, 5, 4, 6, 2, 4, that
is to say that the cylinders 6 and 4 taken into account in FIG. 1
are normally displaced by an angle .beta. of 120.degree. on the
camshaft.
In this case of the present invention, the cams 14D of the starting
cylinders are displaced rearwardly through an angle .alpha. with
respect to the direction of rotation of a pump corresponding to the
arrow B, in order to reduce the injection lead, which is favourable
for starting purposes and for reducing the maximum pressure of the
cycle.
In the case of the cylinder 4, the lead is reduced by an angle
.alpha., for example having a value of 10.degree., as compared to
the injection lead for the cylinder 6.
The same result may be obtained in the case of in-line pumps, by
retaining the conventional camshaft of the equal division type, but
by employing plungers to feed the starting cylinders 3, 4 whereof
the profile 41D illustrated in FIG. 5B offers a reduction in height
x as compared to the profile 25P (FIG. 5A) corresponding to the
piston feeding the power cylinders 1, 2, 5, 6 in such manner as to
reduce the lead and to limit the rate of delivery.
With pistons having a form 26D (FIG. 5C) and comprising a sloping
upper profile 27, the same two actions are obtained in increasing
manner when the flow rate established increases.
In FIGS. 8 and 9 has been illustrated a rotary injection pump
comprising a barrel 28 wherein is rotatably installed a core 29
entrained in rotation by a spindle 30 and having axial bores
wherein are slidably installed plungers 31 which co-operate with a
cam-shaped ring 32. The fuel which is fed in through the pipe 33 is
impelled by the pistons 31 into a passage 34 of a rotary
distributor 35, the said passage 34 opening at 36 into a
distribution head 37 having outflow passages for each cylinder,
such as 38P and 38D, which lead to the corresponding power 6 and
starting 4 cylinders.
In FIG. 9 has been illustrated the positioning of the radial
passages leading to the different cylinders of the 6-cylinder
engine of FIG. 1, and in particular the passage 38P leading to the
cylinder 6 normally staggered by the angle .beta. (120.degree.)
with respect to the passage 38D and corresponding to the
theoretical position 4a. According to the invention however, the
injection lead decrease is obtained by an angular displacement of
the passage 38D in the direction of rotation, by an angle
.alpha..
Analogously, the passage 39D leading to the cylinder 3 is displaced
through an angle .alpha. with respect to the theoretical setting
3a.
It is equally possible to act on the positioning of the cam ring 32
to accomplish the same result.
In FIG. 9 has been illustrated another means rendering it possible
to establish a differentiation between the rates of flow by making
use of non return valves 42D on the passages 38D and 39D of the
starting cylinders 3 and 4, which valves have a high rate of flow
-Q' corresponding to a rate of flow injected Q.sub.o - Q' = q for
the starting cylinders.
On the passages 38P leading to the power cylinders 1, 2, 5, 6 are
installed non return valves 42P having a small volume -q',
corresponding to a rate of flow injected Q.sub.o - q' = Q for the
power cylinders.
In FIG. 6 has been illustrated a V-8 cylinder engine, wherein the
starting cylinders 103 and 104 are paired under the same cylinder
head 109. It should be noted on this subject that the action of the
two cylinders 103 and 104 is not necessarily evenly spread over the
mechanical cycle of the engine.
In this embodiment, the cylinders 103 and 104 have a true
compression ratio increased by reduction of their unswept volume
and/or of the inlet closure delay. These same cylinders have a rate
of injection flow which is reduced and delayed by the same combined
or coupled means set forth in the case of the 6-cylinder
engine.
FIG. 7 is applicable to the case of a flat 12-cylinder engine,
wherein three cylinders 210, 211 and 212 are starting cylinders and
are evenly spaced apart over two revolutions and grouped in such
manner as to concentrate the starting means. Taking the firing
sequence 201, 210, 205, 207, 203, 211, 206, 209, 202, 204, 208 into
account, this corresponds to a constant spacing of 240.degree. over
two revolutions of a four-stroke engine. Operation: the engine is
commonly constructed to withstand a given maximum cycle pressure;
it is of importance that the starting cylinders having a high
compression ratio should receive the same maximum pressure as the
power cylinders having a low compression ratio. By comparison to
the engine operating in conventional manner, whereof all cylinders
work in identical manner with a pressure of the order of 120
kgs/cm.sup.2 for example, endeavours will be made in the case of
the engine embodying selective compressions in accordance with the
invention, to maintain a maximum pressure of the order of
100kgs/cm.sup.2 in both kinds of cylinders, whilst maintaining
performance and improving ease of starting, by an appropriate
injection setting of which an example is given hereunder by way of
information:
Conventional Selective engine compression engine Starting Power
cylinders cylinders
__________________________________________________________________________
Volumetric compression ratio 17 20 13 Inlet delay 45.degree.crank
20.degree. 50.degree. angle Maximum rate of delivery per cycle 60
mm.sup.3 /cycle 40 70 Injection lead 20 crank angle 15.degree.
25.degree. Maximum pressure 120 kgs/cm.sup.2 100 100
__________________________________________________________________________
It is considered that the power rating of a given engine making use
of selective compression may be improved, as well as the ease of
starting, whilst reducing the maximum stresses.
Beyond the advantage of improving durability, it is pointed out
that the power cylinders have the benefit of a large unswept volume
and of a lower compression ratio allowing of an appreciable
reduction of the proportion of nitrogen oxides.
One inventive feature of interest is its ease of adaptation to the
engines actually available, which by means of relatively
uncomplicated and inexpensive modifications renders it possible to
improve their performance and above all their durability.
In FIG. 10 has been illustrated an internal combustion engine
comprising power cylinders 301 and 304 and starting cylinders 305
and 306 which are fed through an inlet manifold 307 from a
supercharger 308. Constrictors 309 and 310 comprising calibrated
orifices 311 and 312 arranged at the inlet ports of the starting
cylinders 305 and 306.
These constrictors 309 and 311 consist in particular of a plate
equipped with a calibrated orifice and situated at the joint with
the flange of the inlet manifold.
* * * * *