CH158334A - Turbo-compressor unit for supercharging internal combustion engines. - Google Patents

Description

Turbo-compressor unit for supercharging internal combustion engines. The object of the present invention is a turbo-compressor for supercharging internal combustion engines. The principle of supercharging: introduction of a greater charge at a pressure higher than atmospheric pressure in an internal combustion engine, is known; . for the propulsion of compressors intended for this purpose, he has tried to use the exhaust gases containing a considerable quantity of heat, pressure and kinetic energy.

The simplest design would be to feed turbine nozzles by connecting each of these nozzles to each of the exhaust ports of a multi-cylinder engine. This is impractical, since a turbine only performs satisfactorily when operating under a constant Pressure and temperature drop, while the Pressure and energy of the exhaust gases is constantly varying. During the escape time. In addition, the need for long and tortuous passages to reach these nozzles causes enormous friction loss, given the high velocity of the exhaust gases. On the other hand, these nozzles, if properly designed, throttle the outlet of the gases increasing the back pressure, which results in a decrease in engine power and other disadvantages.

An immediate and known method therefore consists in using an intermediate tank of dimensions which make it possible to maintain a constant pressure despite the intermittent influx of exhaust gases from the engine. From this tank, the nozzles of the turbine could be fed in a suitable way, its high pressure on the other hand would create a counter-pressure on the engine, not to mention its weight and size.

The turbo-compressor group. forming the object of the invention, overcomes ä. these disadvantages. It is characterized by a device ensuring the supply of the turbine, at constant pressure, by means of an annular chamber in which the exhaust gas circulates with a rapid movement of rotation by playing the air of the gaseous flywheel, and by a rotor comprising a raue which has deug blades, so as to. serve as a turbine driven by the exhaust gases, and on the other as a centrifugal air compressor.

The attached drawing represents, by way of example, one form of execution of the object of the invention, as well as some variants.

fig. 1 is a cross-section of a turbocharger according to the invention; fig. 2 is a side view with a partial section; fig. 3 is a perspective of the blades; fig. 4 is a detail of an oil distribution ring seen from the front; fig. 5 is a view of a casing partially in section; fig. 6 partially shows a geometric aspect of Paubage's construction; Figs. iä. 10 show the blade of the turbine; Figs. 11 a. 15 show an improved construction; fig. 16 is a diagram of an ä motor. non-supercharged four-stroke; fig. 17 is a meme for the running engine of a turbocharger; fig. 18 schematically shows a chamber in the first case of a 12 cyl. V-shaped; fig. 19 shows the application to another engine.

Figs. 1 to 2, it is apparent that the outer parts of the turbo-compressor consist of: a bearing box 2 with a cover 4, an annular section 6 which is called the compressor casing, containing air passages and forming a diffuser transforming the air discharge velocity into pressure, another Part 8 constituting the first cai-ter of the turbine comprising the so-called secondary nozzles supplying the latter and a Part 10 fixed to the engine, in order to receive the gases therefrom exhaust. This Part also comprises the so-called primary nozzles serving to lead the gases from the engine to the velocity equalization chamber formed by this latter Part and the casing of the turbine. These two Parts are suitably bolted together with joints between them as described below. The single rotor 12 bearing the blade of the tubrine and the compressor vanes rotates in a space formed by Parts 2, 6 and B.

Different parts are hollowed out, in order to form Passage for the cooling water coming from the water circulation of the engine. An insulating ring 13 between parts 6 and 8 is intended to insulate parts 6 from excess heat.

In order to obtain good thermal efficiency, compact and simple construction and suitable RTI ID="0002.0275" WI="18" HE="4" LX="1540" LY="748"> cooling, a rotor door-to-door was essential, for reasons that we will see.

A mass rotating at a high speed, such as such a rotor, tends to rotate around its center of mass which cannot be made to coincide exactly with the side of the shaft supporting it.

As the rotor speed increases, the centrifugal forces, due to the eccentricity of the center of mass and tending to deform the shaft, increase in proportion to the square of the angular velocity.

When the shaft reaches a critical speed at which the strong centrifugal force is equally compensated by the elastic resistance of the shaft, the bending and consequent vibrations of the shaft reach a dangerous amplitude: Aug speeds above ä. At critical speed, the rotor rotates around its center of mass, the elastic forces of the shaft being insufficient to force its rotation around the shaft. As the critical speed depends on the shape of said shaft, it is obvious that a good design of the latter allows a rotor to turn at a predetermined speed above or below the critical speed of the rotor and to his tree.

In order to force a rotor of a certain mass turning at high speed to turn below its eritical speed, a very massive shaft is needed. As it is generally impossible -to use a sufficiently massive shaft to resist aug efforts of very Brande speed of a turbine. impulse, an generally uses a "flexible" shaft which allows the rotor to rotate above its critical speed.

Such a flexible shaft is not applicable to a cantilever rotor, because when its speed approaches the critical speed during acceleration, the vibrations of the cantilever rotor would reach a dangerous amplitude capable of destroying the tree.

The bearing construction which will be described eliminates the critical speed by suppressing the forces determining this speed, ie the elastic forces tending to. to prevent the rotation of the rofor around its center of mass. In other words, even at low speed, the rotor will turn around its center of mass, so that it will have no tendency to vibrate when accelerating the fix. 1 explains the construction allowing this. The shaft 14 which preferably forms an inUgrant part of the rotor (although it may be fixed to the latter) is provided with sleeves 16 and 18; the latter being able to rotate freely with respect to the shaft. Between a shoulder on the sleeve 16 and a nut screwed on the shaft is fitted the inner bearing of a ball bearing (or roller bearing of the self-centering type at Borge, so as to obtain the centering, the outer bearing being therefore fixed to the bearing shell.

A series of sleeves 22 fitting together and having a slight play between them and the first inner sleeve 18 which they surround, allow free rotation and a slight transverse play. A clearance is also provided between these sleeves and the collars as shown in the drawing.

The outer sleeve rotates freely in a sleeve 24 fitted to a sleeve 30 integral with the housing 2. The sleeve 24, as indicated in detail in the fix. 5, consists of an outer cylindrical portion 24b and a thin annular radial section 24e. Through the junction of 24e and 24b a number of fairly wide grooves 24d are cut in 2413. The remaining portion between the grooves is recessed on the outside as shown.

in 24th, thus forming a number of laues of springs supporting the entire pad and allowing soll adjustment Jans a radial direction, thereby allowing a slight dPpl: the angular cement of the shaft.

An oil passage 26 in the shaft receives oil via 27 and a chamber ?g is supplied with pressurized oil from the normal oil circulation of the engine. Radial orifices in the shaft and the sleeves 18 and 22 allowing the oil to pass through the chamber surrounding the nested niancho-is, the first passage of the oil from this chamber to the ball bearing and to the compressor of air being stopped by centrifugal action rings 31 cooperating with fixed deflectors directed towards the interior. Oil exits the chamber through port 32 through 25 and finally at 34 is directed to the crankcase of the engine. Oil distribution between the rings is provided by grooves in the outer surface of the sleeves 18 and 22, the sleeve 18 also having grooves on its inner face adjacent the shaft. As in,li qu6 in 1a. fix. 4, the first side of each groove presenting itself first to the direction of rotation, offers a much greater angle with the cylindrical surface of the sleeve than that made by the other side, so that a crushing effect or Wedging occurs, ending a positive oil pressure in the space between the sleeves. Ei) consequence, this space between the inanchons is filled by a film of oil in iiion vement continuously renewed.

A small amount of oil is delivered to the ball (or roller) bearing 20 by a wick 36, pressed by a spring and communicating with the oil chamber 28. Excessive lubrication of such a bearing should be avoided.

It will be seen that angular displacement of the shaft around the bearing 20 can occur freely, this being of the self-centering or deep-bore type and that a limited movement of the shaft transversely with respect to Axis movement can take place through the play of the sleeves 22 and the flexing of the sleeve 24, oil being introduced or extracted therefrom by such movement. Consequently, almost at the same time as the rotation begins, a slight displacement of the rotor and of the shaft, in order to bring the center of mass back into the axis of rotation, can occur and the eccentric movement of the shaft can take place by displacements radials of the sleeves accompanied by a change in the thickness of the oil films. At the same time, the damping action of the oil stops any vibration at its origin, its viscosity opposing a rapid change in thickness of the films between the sleeves.

As a result of the lack of shaft resistance, there is no longer a critical speed at which dangerous vibrations could occur and the shaft need only be of sufficient strength to support the dead weight of the rotor. The use of one of these bearings, preferably adjacent to the rotor, is sufficient to completely eliminate any vibrations at high speed.

Turning to particulars relating to the drive of the turbine, with reference to FIGS. 2 and 3, showing that the inner section 10 of the shroud is provided with fittings enabling it to be fastened to the exhaust side of an internal combustion engine, so as to take the place of the normal exhaust chapel, 1a. constructian dkrite applies <B>ä</B> a six-cylinder in-line four-stroke engine. Accordingly, there are six passages 38 aligned with the exhaust passage of the cylinders, in order to direct the exhaust gases to the primary nozzles 40 communicating with an annular chamber 42 formed by the parts 8 and 10. adjusting. While the passes are formed by two Cooperating Parts 8 and 10, the sections are indicated as dune Beule pike in fig. 3, this for clarity. The annular chamber 42 is designed in such a way that the gases which enter can circulate around it with a minimum loss of speed. For this, the walls are polished, curves as much as possible of Brand radius are established; direction valves can be interposed in order to avoid eddies and to present passages of suitable dimensions and shapes. In addition, the nozzles are arranged in such a way as to discharge the gases in a direction coinciding as far as possible with the tangent to the di- ="1200" LY="502"> rection of the current next to their orifice in order to avoid interference from. carbonated fillets.

From the chamber 42 extend nozzles 44 (Fig. 1) intended to produce a further transformation of heat into kinetic energy and to direct the gases already moving at high speed towards the blades 46 of the turbine. It will be noted that these nozzles preferably formed as passalts in the outer part 8 are arranged tangentially to the path of the gases arriving in the chamber 42 and in line with the discharge nozzles 40. In order to ensure the best efficiency, as many nozzles secondary nozzles as possible should be in a straight line with the primary nozzles, in other words, each gas jet must leave the corresponding Primary nozzle in such a direction that it can pass through some of the secondary nozzles. @ without being subjected to a Brand of deflection and thus to bring the gases to the secondary nozzles with the Brande possible speed. Thus the gases enter the secondary nozzles at the very high discharge velocity of the nozzles 40 and this initial velocity increasing during the adiabatic expansion in the nozzles 44, there results a very high discharge velocity in the blade of the turbine. . It will be seen that the velocities resulting from the two pressure drops in the primary and secondary nozzles are directly added in order to obtain the maximum kinetic energy of discharge in the blading.

In order to obtain an approximately straight direction of each jet of gas before proper entry into the rotor, the secondary nozzles 44 are shaped so as to form portions of an annular space bounded by two hyperboloidal surfaces of rotation. Deloading from the secondary nozzles takes place in a continuous fashion, while deloading from the primary nozzles is intermittent. The flow velocity through the Primary nozzle corresponding to a given cylinder is very high immediately after the opening of the corresponding exhaust valve, falls rapidly During the time of the exhaust, no flow takes place. For the next three strokes (if it is a four-stroke engine). The secondary nozzles are much narrower.; that the primary nozzles, so that the gases Passing, in the chamber 42, during the opening of a valve, cannot flow completely immediately through the secondary nozzles. The egces will flow around the chamber 42 by rotating in it, operating ä. like a "gasoline flywheel", the energy remaining kinetic and not being transformed into heat or pressure. Then, when the secondary nozzles can absorb it, this excess gas will flow through these secondary nozzles into the blading. As a result, with proper proportions, a continuous discharge will be obtained at nearly constant velocity through the secondary nozzles in the turbine blade. Since the energy of the gases in aegean is maintained in an einetic form, there is no increase in back pressure, very high counter pressure with the intermediate gas reservoirs employed to date. On the contrary, there is a secondary action increasing the efficiency of the motor. As indicated, the velocity of the escape decreases very rapidly, during the last parts of the escape time; when the velocity in the Primary nozzle becomes less than the velocity of the gases circulating in the chamber 42, Ges. gases generate an ejector effect as they sweep tangentially across the nozzle opening, thereby reducing the back pressure in the cylinder.

This latter effect would make the use of chamber 42 advantageous even without a turbine. With properly designed primary and secondary nozzles, the back pressure when the valves open is no more Brande than Gelle which would occur if the engine were operating normally, given the friction in the openings. exhaust and valve openings or restricted orifices. At the same time exhaust through the secondary nozzles takes place continuously eliminating the intermittent exhaust noise that normally causes RTI ID="0005.0275" WI="13" HE="4"LX ="1204" LY="452"> the use of a silencer. Consequently, it is clear that the use of such a device as a silencer, instead of increasing the back pressure, decreases it.

The construction of this annular chamber can differ radically from the Gelle described, but it depends on the type of engine to which it is applied and in particular on the exhaust ports of the latter. The one described is suitable for a normal automobile engine with overhead valves and six cylinders in line. If the engine and the turbocharger were built and designed as a unit, the exhaust valves could take a needle shape, in order to directly close the primary nozzles, or alternatively the valves, while being seated in their usual place, could be constructed so that their seats also serve as primary nozzles. With a two-stroke engine, the primary nozzles would coincide with exhaust ports in the cylinder walls , uncovered when the first piston approaches or reaches the first ba.s of its stroke.

fig. 18 schematically shows the arrangement of the annular chamber in the gas of an engine. twelve cylinders in V. Width of the turbocharger rotor coinciding with an bisector of the V. The primary and secondary nozzles being designated by <I>A</I> and <I>B.</I>

Figs. 7 to 10 show that turbine blade 46 is shaped as semi-circular indentations on the right side of rotor disc 12. FIG. 7 clearly indicates the operations of their formation. An annular groove 48 is first cut on 1a. side of the disc. Then semi-circular notches 50 are cut in the face, transverse to the groove 48, extending inward and forward in the direction of rotation. Since the adjacent notches 50 intersect near their midpoints, when they converge internally in the exposed modification, the surface of the. groove 48, as indicated in the figures, disappears. An annular protuberance 52 (Fig. 1) engages in the groove 48 inside the nozzles 44, so that the gases from the nozzles enter tangentially as indicated in the direction of rotation in the outer parts. notches 50 and are directed inward and rearward around the protuberance 52 and discharged into the exhaust passages 54 communicating with a chamber or vaute having its exhaust 56 connected to a final exhaust pipe - payment. Inside the exhaust chamber is an Erous Part 58 communicating with the atmosphere; provided with a mouth flaring and brought closer to the rotor at 60. The gases discharged from the blades exert an ejection effect at 60, thus sucking air through 58 which licks it. face of the rotor and the 1st cools.

Such a blading will withstand the highest temperatures and will have a good performance. We will now design a blading of this improved type with optimum performance.

In the first type described above, the lack of guidance of the gases at the exit of the dawn, resulting from the absence of any partition, lowers the efficiency; the second type which will be described is an improved construction remedying this defect, while retaining the advantages of the preceding type. Cut cavities could be placed at 1a. periphery of the rotor, so that the inlet and outlet sections move at the same linear speed. It has been found that by using similarly shaped vanes in a radial direction with the gas inlet as outboard as possible, such that the peripheral velocity at the inlet is greater than at the outlet. , the yield was improved. The following table gives the efficiencies obtained as a function of the different ratios between the peripheral speeds of the outlet and those of the inlet of the blade, assuming the eapacite of the equal turbine, as well as the same losses due to, the. friction and shock:

yield ratio 0.50 84.0 0.75 78.2 1 73.4 2 55.8 Fig. 11 which shows a two-stage turbocharger, and FIGS. 12, 13, 14 and 15 illustrate RTI ID="0006.0271" WI="8" HE="4" LX="1510" LY="828"> this improved blade construction. The inwardly converging notches constitute the vanes 34; they consist of the intake 112 and exhaust 114 parts (fig. 15), the partitions between the blades being hollowed out as indicated at 116 (by cutting into the face of the rotor, before any other operation); to receive an annular protrusion 117 separating the intake nozzle from the exhaust passage. In conjunction with the vane bulkheads, this protuberance forms the usual curved Passage through which the gases entered are diverted and from which they escape.

It has been found that for the radially errous type of blading with the inward gas flow opposing the centrifugal forces, the efficiency can be increased by making the exit angle smaller than the entry angle. In order to achieve this result, the part of the partition below the recess 116 is brought from its position 120 indicated in dotted line (fig. 13) to position 118, in which it is welded, at 122, as clearly indicated by fig. 15. Note that as the load angle decreases, the height of the outlet is increased to achieve the proper gas exhaust area. The result is a lighter, more compact rotor construction. The vanes are designed to complete the expansion as indicated above. In the previous construction (First type) only one milling is necessary to constitute each vane, but as a result by providing a Maximum number of vanes, the partitions between the Discharge Parts disappear, from which loss of suitable guidance. In this latter construction (second type), the Separation walls are of uniform thickness over the entire radial extension of the vane and consequently the vanes are outwardly converging. This complicates a little bit their arrangement; But complete and proper guidance of the discharge is achieved. This construction is facilitated by the fact of forming the partitions first in plan, then by folding the inner part as indicated, but it is obvious that an operation of modified size would allow a construction without this folding.

The advantages of this type of blading described in two variants are clear, if one considers the disadvantages of the usual peripheral blading. While such peripheral blades suitably fixed to the rotor are satisfactory at a relatively low temperature of a steam turbine, they can hardly resist satisfactorily at the same time. combination of higher gas turbine speed and temperature. The reason is obvious: to obtain a certain deflection of the working fluid, a certain active concave Faubage surface must be provided. This surface is subjected to the heat of the fluid.

On the other side, a relatively small area of contact between the blades and the rotor disk resulting from the peripheral construction must allow the flow of the heat received by the blading, this area being generally considerably less. at the active surface.

In the construction described, all the active surfaces are, at the same time, surfaces capable of conducting heat to the rotor disc. In addition, the blades cannot burst centrifugally and the dead weight added to the disk periphery is minimal. The shape of the blades is suitable for. give Maximum yield. Moreover, due to the groove .1S (fig. 1), the danger of trog Brandes stains for the parts remote from the main mass of the disc is karte.

The ad, duction of heat to the rotor causes evideminent the need to refroldir it. Cn saw the action of Part 60 to cool the face of the rotor. We will see later that the same rotor serves as the rotor for the centrifugal air compressor and RTI ID="0007.0275" WI="7" HE="4" LX="1441" LY="516"> that it is swept by Fair Fresh. The advantages of the centrifugal compressor which will be described will be made more evident by recalling the disadvantages inherent in current centrifugal compressors. These Compressors have vanes which give the plane form extending radially to the axis of rotation, have surfaces having the elements in straight lines parallel to the axis, the sections by planes perpendicular to the axis are curved in such a way that the exit angles at the Periphery are less than right angles. These two forms have (b- , serious disadvantages.

First: the flow entering through the gateways is subject to two sudden changes in flow direction; that occurring on entry when the current changes from its axial direction to a direction matching the rotation of the rotor, the direction changing to a direction radial towards the periphery. Stationary guides are sometimes used to guide incoming air; But this increases the yield only very slightly, because of the beaded bands out of place by Continuation of deflection.

The second serious disadvantage is that such fins, of the open radial type, must be fixed directly to the hub, even for reduced speeds, it is impossible to fix auxiliary guide fins, in order to maintain an adequate conrant between the Outer parts of the passaltes. In general, the air gaps are closed on both sides by the stationary walls of the casing, which requires clearance resulting in significant losses. These disadvantages already occur at low speed; other additional factors make such compressors difficult to use at very high speeds.

The fins have, compared to ä. their length, a very curved attachment line with the nucleus. At high speed, curved fins are subject to enormous centrifugal forces tending to bring them back into a radial direction.

The construction indicated. by Figs. 6 and 7, reduces or eliminates these inconveniences, ensuring straight lines to the air current from the inlet to the periphery, as well as the fins having radial elements eliminating any deformation due to centrifugal forces. In addition to the main fins, short auxiliary fins are provided, ast to help maintain a suitable inflow between the peripheral portions of the main fins.

The surface of the rotor 12 (FIG. 1) opposite the casing of the turbine, the outer surface of the side of the shaft, presents on its surface 60 a part of a hyperboloid of revolution, the part of the hyperboloid used being the continuation of the surface indicated by line 60a, that is to say that the part used is entirely on one side of the groove of the hyperboloid. We know that a hyperboloid of revolution is not only a surface formed by the rotation of a hyperbola around its small height, but also the locus of a straight line revolving around an height with which the height has no plane. known, so that the hyperboloidal surface can be conceived as formed by straight line elements. It follows reciprocally that, Aant gives a Series of elements having certain directions relative to an alternation of rotation, a hyperboloidal surface can be constructed. It is this fact that is utilized in designing the surface as discussed below. - grooves for increasing the main fins, similar but shorter intermediate grooves are provided for the auxiliary fins.

The construction lines of fig. 6 clearly show the geometric construction of a fin. As indicated, a series of lines 66 can be constructed passing through the points of the center line of groove 62a and located perpendicular to the rotor 12. ="7" HE="4" LX="1674" LY="542"> can stop aug lines 68 and. 69 where it intersects a conical surface 70 (fig. 7) and a cylindrical surface 72. This elementary surface is in reality a conicohelical surface of variable pitch, although it approximates a plane by which it could be replaced. Each main fin 74 may be constructed on the elementary surface 61 so that the center of gravity of each section by a plane transverse to the rotational page lies in this surface. In other words, this elementary surface is between the faces of the fins.

Each auxiliary fin 76 can be considered as constructed by an analogous elementary surface passing through the grooves 64.

These fins can be engaged in the grooves and welded along their edges; other means of attachment may be provided, among other things they may come from forste with the hub or the air passages may be constructed by cutting them into the mass of the rotor.

The operation is as follows: Air is taken through openings 78 (Fig. 1), into a suction nozzle or volute, inlet 80, from where it passes through the air passages between the fins and is discharges at high velocity into the passes of diffuser 82, where part of this high velocity is converted into pressure. From the vents of the diffuser air is charged into a volute 84 terminating at 86 in a suitable conduit to the carburettor, from where the detonating mixture is introduced into the engine; The carburettor could be placed upstream of the compressor. For a Diesel or any other injection engine, there is obviously no carburettor. The diffuser 82 is cooled by the fins 83 bathed in the cooling water of a circulation connected to the Gelle of the engine.

The goals sought by this compressor construction are obvious. The direction of current through an air gap adjacent to surface 60 (FIG. 7) will be along straight lines which strike the elements of the helical surface. Knowing the average speed of rotation and the quantity of aii delivered under normal conditions, the direction that an element in a straight line should have, in order to "draw" air into chamber 80 (fig. 1) and pass without collision and at an angle suitable for the passages of the outlet diffusers 82, can be easily dAerminee. Knowing this, the shape of the hyperboloid surface and the fins follows. For proper design, air enters the Aales in the direction of their smooth surface, runs in a straight line along the surface 60, and is also smoothly discharged into the smoothly arranged passages 87. receive Fair without shock. The exit angle of the fins is less than 90 (approximately 65), thus ensuring maximum efficiency. While the air adjacent to the surface 60 is flowing exactly in a straight line, all that passing through the air passages is flowing close to the straight line, so that there are no losses through the air. bumps and gels due to manufacturing irregularities are reduced to a minimum. The peripheral sides of the passages are preferably formed by the conical surface 88 of the box 2, although if necessary an annular partition may be carried by the outer edges of the fins.

The mechanical advantages of this construction are also important. The line of attachment of the fins to the disc is very long, compared to the first surface, thus ensuring maximum solidity. For this, and because of the curvature of the surface 60, it is possible to place secondary guide fins 76 between the main fins, so as to guide the current and avoid eddies due to the fact that the divergence of the main fins becomes too brande for good guidance without] the auxiliary fins.

The fact that the vanes are attached along straight lines of the rotor makes them easy to manufacture allowing them to be brought in from the bridge with the disk.

The elements of the fins being radial, the centrifugal forces are exerted in tension only, both on the fins and at the first junction with the rotor, this RTI ID="0009.0274" WI="5" HE="4" LX ="1410" LY="490"> which gives them a solidity making them able to resist speeds. of the order of 60,000 revolutions at mid-speed.

Many changes can be made to the described compressor. while remaining within the scope of the invention.

For example, a construction similar to Gelle described having the radial fin elements, results in a slight separation of a straight flow through the outer portion of the passages. An alternative construction in which the walls of the fins are straight lines in the direction of the scale could be adopted; in this Gas, the fin element would deviate slightly from a strictly radial direction, although not enough to allow destructive forces to appear.

*The fins, as well as the face of the disc in contact with a fast air current, contribute to the cooling of the rotor.

In order to prevent the assage from building up pressure in the exhaust gases entering the turbine housing, flexible segments 90 form a labyrinth seal; As the annular assage 92 is inclined in the opposite direction, a Pressure is created in this annular space, corresponding to the velocity of the gas jet, in order to balance the Air Pressure and thus contribute to. avoid a serious leak.

From the detailed description above, it follows that the turbo-compressor unit described comprises various new constructions of details. Their combination into a new type. of turbocharger ä. high efficiency is particularly advantageous, since in addition to this efficiency, the aggregate is simple and compact and eliminates the disadvantages of long pipes connecting it to the engine, the engine-turbo-compressor group itself being compact.

As indicated above, the. The construction described is intended for use in connection with an inline six cylinder engine. For a v-shaped engine, figure 19 shows the particularly compact arrangement of the turbocharger inside the V, the primary and secondary nozzles being indicated by l1 and <I>B.</I>

fig. 11 shows a device with two stage compression for very high altitudes, 10 representing a cantilever disc. as already described, having a turbine blade 34 on one side and compression vanes 14 on the other; a second disc 12 on the same shaft carries the compressor vanes 16 and constitutes the second stage compressor; an elastic pad 8 like the one described is provided between the two discs. THE. two compressors are, as already said, connected in series, in order to ensure a high pressure; air coolers are provided (18) and are arranged in a compact manner. Other layouts may be used.

The theory of overfeeding is known, as well as its practical advantages and disadvantages. The type of turbo-compressor described, thanks to the "gaseous flywheel" device, considerably reduces the back pressure and uses the greatest part of the energy of the exhaust gases, results which could not have been reached practically before. The high efficiency of this group results in a strong increase in the power of the internal combustion engine, to which said device is applied.

fig. 16 is the known diagram of a new supercharged four-stroke internal combustion engine, the atmospheric pressure being represented by the line -A-t. fig. 17 is the diagram of the same engine running the turbo-compressor described. The Pressure obtained by the compressor is represented by the line S-u. The positive surface J is considerably increased with respect to FIG. positive surface can be obtained and it is seen that, thanks to the device of the "gas flywheel", most of the energy contained in the exhaust gases and represented by the surface n is used in these turbines, whereas hitherto, when using only an ordinary intermediate reservoir, only a small part of this energy could be used.

The large increase in power RTI ID="0010.0274" WI="4" HE="4" LX="1862" LY="588"> resulting from the use of supercharging allows for more power per kg. engine; hitherto this increase was limited by the energy absorbed by the compressor. The turbo-compressor unit described makes it possible to obtain very high supercharging pressures, by using stage compressors, this without increasing the counter-pressure of the exhaust gases; hence an increase in power of the new turbo-compressor engine assembly achieved to date.

Claims (6)

CLAIM Turbo-compressor unit for supercharging an internal combustion engine, characterized by a device ensuring the supply of the turbine, <B>ä</B> Constant pressure, by means of an annular chamber, in where the exhaust gas circulates with a rapid rotational movement playing the role of gaseous flywheel, and by a roter comprising a disc which has two blades so as to serve on one side as a turbine driven by the exhaust gases, and on the other side, centrifugal air compressor. SUB-CLAIMS 1 Turbo-compressor unit according to claim, further characterized in that the turbine blade ä. gas is constituted by semi-circular notches made in the first disc and disposed radially and angularly with respect to. the face thereof, and in that the centrifugal compressor vane is formed by fins extending radially from straight lines which are elements of a hyperboloidal surface of the disc. 2 Turbo-compressor unit according to sub-claim 1, characterized in that the first centrifugal compressor is provided with augi.liaires fins provided between the outer parts diver gentes .des main fins. 3 Turbo-compressor unit according to claim, further characterized in that the rotor shaft is cantilevered on at least one resilient bearing. 4 Turbo-compressor unit according to sub-claim 3, characterized in that the first elastic bearing is formed by sleeves embo "ztes separated from each other by a film of oil circulating between them, the sleeves being carried by parts forming a spring, a socket. 5 Turbo-compressor unit according to sub-claim 1, characterized in that the partitions between adjacent blades of the gas turbine 5 are bent in such a way that the exit angles of the gases with respect to the corresponding face of the rotor are smaller. than the entry angles. 6 Turbo-compressor unit according to sub-claim 3, characterized by a pipe bringing cooling Air towards the face of the rotor serving as a turbine.