JPH0366500B2 - - Google Patents

Description

TECHNICAL FIELD The present invention relates to a bearing support structure that supports rotation of a compressor and a turbine of a turbocharger.

Context Reference for Related Applications This application is a CIP application of US patent application Ser.

BACKGROUND OF THE INVENTION Turbocharging is a means of greatly expanding the power range and flexibility of internal combustion engines, and has become an accepted technology in practice, often with 200
It has become necessary for powerful diesel engines with horsepower or more. Turbocharging is also used, for example, in aircraft engines to maintain power as altitude increases.

Effective applications of turbocharging to internal combustion engines typically increase power output by 50-100 percent,
Reduces full load specific fuel economy (SFC) by 5-10 percent. The decrease in specific fuel consumption is attributable to two items. First, at a given speed, the friction of the internal combustion engine remains relatively constant even though the power output increases considerably. This results in an effective improvement in mechanical efficiency. Second, if the efficiency of the turbocharger components is high enough and the exhaust temperature of the engine in which the turbocharger is used is high enough, there will be positive pumping work added to the net cycle output. Become. In a four-stroke internal combustion engine, as is clear from its acupressure diagram, it is clear that negative pump work, or pump loss, generally occurs. Page 223 (1) Pump Loss Section and Edward F. Obert, Intext Educational Publishers
“Internal Combustion Engines and Air Pollution” published by Educational Publishers.
Air Pollution”, page 156). However, in the case of a turbocharger that satisfies the above conditions, positive pumping work will instead occur. ] Turbocharging is generally considered a means of reducing full load specific fuel economy and increasing horsepower. Organizations currently using turbochargers,
For example, with the development of large powerful diesel engines, aircraft engines, competitive automobile engines, and similar engines, the foregoing understanding is valid, and in limited applications these engines can be used for most of the work cycle. It is operated at or near full load for a period of time.
However, in many applications, it is not necessary for these engines to be operated at or near full load for extended periods of time. In many practical applications, these engines typically operate at less than 50% power, and in many applications these engines operate at less than 20% power during most of their operation. be done.
Examples of these applications are automobiles, light or medium trucks,
Engines used in generator sets, compressors, tractors, building equipment, and similar equipment.

Engines operating in these low power environments are very inefficient. In diesel engines,
This inefficiency is the result of a decline in thermal efficiency as combustion temperature decreases. This is because friction in an internal combustion engine remains relatively constant regardless of load. In gasoline engines, this inefficiency results from increased pumping losses as the load decreases. This is because friction in an internal combustion engine remains relatively constant regardless of load.

Therefore, by using smaller and more efficient engine dimensions, i.e. engines with smaller displacement ratios (1/2 displacement x engine speed product for four-stroke engines). Part-load fuel consumption can therefore be improved by reducing displacement, reducing operating speed, or a combination of both. In gasoline engines,
This improvement occurs as a result of reduced engine friction and reduced pumping losses. In diesel engines, this improvement is the result of higher thermal efficiency and reduced engine friction due to higher combustion temperatures. In many applications, turbocharging may be used to allow the use of engines with smaller effective engine dimensions. By turbocharging, it is easy to obtain twice the naturally aspirated power per cubic inch of displacement, and in some cases triple the naturally aspirated power per cubic inch of displacement.
However, attempts to turbocharge smaller engines have generally been unsuccessful. This failure is caused by journal bearings and flat disc type bearings.
This is due to current designs in which the turbocharger is assembled around the thrust bearing. This type of bearing system requires 1 to 3 horsepower (depending on the particular turbocharger and speed required in the application) just to overcome friction. This loss is not significant in applications where the turbine is required to produce more than 30 horsepower during the compression process (typically 200 horsepower or more engines), but when turbocharging engines less than 100 horsepower time becomes very important. For example, if the turbocharger turbine power is 60-80 horsepower, bearing friction losses of 2-3 horsepower are not significant. However, in a smaller turbocharger where the turbine power is only 15 hp, bearing friction losses of 2-3 hp or even 4-5 hp due to the higher rpm at which the smaller turbocharger is operated are: This represents almost 1/3 of the total turbine horsepower generated, a totally unacceptable loss.

Also, currently used bearing systems require significant radial and axial clearance to allow oil flow and rotor stability. These clearances are translated into relatively large clearances associated with the blading of the compressor and turbine rotors, thereby affecting the efficiency of both the compressor and the turbine. For example, journal bearings and flat plate thrust bearings commonly used in modern turbochargers require a clearance of 0.015 inch between the turbine and compressor blades and the surrounding structural members. If the height of the blade is 1 inch, the ratio of the clearance to the height of the blade is 1-1/
Only 2%. However, if a smaller turbocharger is desired, with a vane height of, say, 0.2 inches, a 0.015 inch clearance between the vane and its surrounding structural members would be 7-1/2 of the vane height. Become a percentage. Therefore, in larger turbocharger applications 0.015
Where inches of clearance may be acceptable, it becomes completely unacceptable when smaller turbochargers are designed. Therefore, in smaller turbochargers, this clearance becomes increasingly critical to the overall performance of the turbocharger and ultimately to the performance of the engine.

Many bearing failures are the result of insufficient engine oil pressure or contamination of engine oil during startup. When high speed journal bearings are used in conventional turbochargers, continuous oil flow is essentially required to stabilize the shaft and carry away the heat generated by viscous friction. Oil flow is also required to carry away the heat transferred to the bearing system from the adjacent turbine (which operates at temperatures as high as 1600°). In conventional turbochargers, even if anti-friction ball bearings are used in place of journal bearings, a continuous flow of oil is required to carry away the heat transferred from the turbine. Therefore, although modern turbochargers require lubrication to operate properly, lubrication is also the cause of many failures. Furthermore, continuous oil flow lubrication requires substantial plumbing and associated structures to supply lubricant to the bearings.

Current turbochargers cannot effectively control the flow of propellant gas through the turbine. There are currently basically two methods used to control the power output of a turbine. The first
The method is to carefully dimension the turbine and turbine nozzle so that at maximum engine operating speed and load, the desired boost pressure does not exceed predetermined limits. The disadvantage of this method is
The available boost pressure at low engine speeds is limited and the response to demand is slow.
The second method used to control boost pressure through the turbine is to size the turbine nozzle and wastegate to produce excess turbine power at maximum engine speed and load. is to use. In this method, when a predetermined boost pressure is reached, an exhaust relief valve opens to bypass a portion of the exhaust gas. Although this method increases available boost pressure at lower engine speeds and improves response to demand, it is completely inefficient in that bypassed high pressure exhaust is simply wasted at the expense of increased engine back pressure. It is. Furthermore, at part load, when the turbocharger is essentially not operating, the small nozzle area acts as a restraint on the exhaust, causing increased pumping losses.

Therefore, a need has arisen for a turbocharger that can effectively turbocharge both large and small internal combustion engines. A need has arisen for a turbocharger with a bearing system that eliminates the problems previously experienced with bearings lubricated by a continuous flow of oil and that makes the most efficient use of the propellant gas that drives the turbocharger turbine. Furthermore, the bearing assemblies supporting the compressor and turbine rotors must readily reduce the desired compressor and turbine rotor clearances.

Conventional compressor housing construction and design for turbochargers also presents a significant variety of problems. Centrifugal compressors are one of the most widely used dynamic compressors in turbochargers.
In this type of compressor, air or an air-fuel mixture enters the compressor inlet, flows toward the compressor rotor, and is accelerated to near the speed of sound at right angles to the inlet flow path. The increase in air pressure is accomplished by reducing the velocity of the accelerated gas discharged from the tips of the compression rotor blades. This process, known as diffusion, is achieved more effectively by reducing the velocity of the gas without creating turbulence, so that a large percentage of the velocity energy is converted to pressure energy and the static Increase pressure.

To facilitate this diffusion process, turbochargers using centrifugal compressors typically have compressor rotor walls that closely follow the contour of the rotor blades from their leading edge to their outer tips. We are prepared. This compressor rotor wall then extends past the outer tips of the rotor blades, forming one of the two walls of the diffuser, and then terminating to form a peripheral gap. The compressed gas flows through this peripheral gap into a peripheral chamber leading to the engine's intake manifold. This wall, which faces the compressor rotor blades, contours closely to the compressor rotor blades, and then extends outward, allows the compressed gas to enter the surrounding chamber after leaving the compressor rotor blades and leading to the engine. Before, the velocity of the compressed gas is reduced evenly. This wall structure therefore greatly increases the static pressure generated by the compressor.

To form this structure, many turbocharger compressor housings were sand-cast such that the outer perimeter housing of the compressor and the compressor wall were one casting member. Usually this is done by using a sand core to form a peripheral chamber leading to the engine's suction manifold.
completed by using. After casting, the sand core is removed to form a peripheral chamber on the wall opposite the compressor, in which gas flows away from the tip of the compressor rotor.

Although die-casting of the compressor housing is substantially less expensive and more precise than sand-casting, it is useful to form the diffuser wall as well as to form the peripheral chamber through which the compressed gas flows to the engine suction manifold. Due to the unavailability of die casting molds, it has not been possible to use die castings of optimal design. It has not been possible to die cast a compressor housing of optimal design because the variable area chamber is necessarily larger than the inlet gap through which gas is injected from the compressor rotor blades. This is because it is not possible to design a mold that can form such a passageway behind the wall facing the compressor rotor blades.

If a die-cast compressor housing is used, the walls normally formed in sand-cast compressor housings are simply removed so that the molds can be combined and split to make the casting. However, without this wall, the gas accelerated by the compressor rotor would be prematurely pumped from the differential user into the peripheral chamber leading to the engine suction manifold. As a result, compressor housings of such construction substantially reduce compressor efficiency and, therefore, compressor performance.

DISCLOSURE OF THE INVENTION In accordance with the present invention, an anti-friction ball bearing assembly is used to support compressor and turbine shafts. Conventional turbocharger bearing systems, which use a combination of journal bearings and flat plate thrust bearings, require significant radial and axial clearance to allow oil flow and rotor stability. I need. This results in large clearances associated with the blading of the compressor and turbine rotors, which affects the efficiency of the compressor and turbine.
However, by using ball bearings, 10 to 15% of the radial and axial clearance or travel of conventional turbocharger bearing arrangements such as
Since it only requires a small amount of movement, the movement can be controlled to be even smaller. This additional control over the movement of the rotating bearing assembly improves compressor and turbine efficiency ratings by allowing reduced blade tip clearance. Furthermore, the use of anti-friction ball bearings improves the engine's specific fuel efficiency by reducing the required turbine work.

In accordance with the present invention, a bearing assembly includes first and second inner races formed on the shaft of a compressor and turbine. A fixed outer race corresponding to the first inner race is attached to the turbocharger housing, and a plurality of balls are connected to the fixed outer race and the first inner race.
Accepted between races. A second outer race ring is provided and slidable relative to the first outer race. The second outer race ring is slidable relative to the turbocharger housing, and a compression spring acts between the turbocharger housing and the second outer race ring to move the outer race ring away from the first outer race. It is biased away and associated with a ball located between the second outer race ring and the second inner race. At the same time,
A first outer race attached to the turbocharger housing is associated with a ball between it and a first inner race on the shaft of the turbine and compressor.

In the present invention, the first inner race 206 and
Because the second inner race 210 is formed around the shaft 160, that is, because both inner races are integrally formed with the shaft, the following advantages are obtained.

(i) It allows the use of larger diameter shafts than if the inner race were not formed integrally with the shaft, and the shaft can be made very rigid.

(ii) Since the eccentricity existing between the inner race and the shaft itself is eliminated, the concentricity between the shaft and the ball bearing can be improved.

(iii) The peripheral ramp or slinger can be located very close to the ball bearings so that the lubricant spray can be directed directly onto the ball bearings.

(iv) Lubrication requirements are reduced because the distance from the axis to the ball center is shorter and the ball bearing can rotate at a slower speed than when the inner race is fitted onto the shaft for the same shaft diameter. and can reduce bearing wear.

BEST MODE FOR CARRYING OUT THE INVENTION FIG. 1 is a perspective view of a turbocharger 20 equipped with a bearing support structure of the present invention. The turbocharger has an external structure 22 consisting of a compressor housing 24 connected to a turbine housing 26 by a V-shaped clamp band 28.

As shown in FIGS. 1 and 2, the compressor housing 24 includes a cylindrical inlet hole 40 having a lateral wall 42 attached to one end thereof and extending outwardly therefrom. A circumferential chamber 44 is attached to wall 42. The inlet hole 40 is the air inlet 5 of the compressor.
0 and the chamber 44 forms the compressor outlet 52. Turbine housing 26 defines a turbine air intake 54 and a turbine exhaust 56 .

In use of the turbocharger, air is drawn into the inlet 50 and compressed air is discharged through the exhaust port 52 to the internal combustion engine to which the turbocharger is attached. Exhaust air coming from the engine is directed to a turbine air intake 54 to drive a turbocharger turbine and is exhausted through a turbine exhaust 56.

A bearing support tube 60 is mounted within the inlet hole 40 by a plurality of vanes 62 extending from an inner wall surface 64 of the hole. A cap 66 is attached onto the end of this bearing support cylinder 60. A piston-type actuator 80 is mounted to the turbine housing 26 by a bracket 82. The actuator 80 includes a controller 84 that is operated to extend and retract a control rod 86 as described below. In order to operate the control rod 86, air is sent to the controller 84 through air pipes 88 and 89 as needed. A sump cover plate 90 is attached to the compressor housing 24 by a plurality of screws 92.

As shown in FIG. 2, the compressor rear wall 100 and the turbine rear wall 102 are connected to the compressor housing 24.
and the turbine housing 26, when these two housings are assembled. These four components are mutually guided and assembled by a V-shaped clamp band 28. As can be seen from a closer look at FIG. 2, only a single V-clamp is required to hold the entire assembly together. As previously mentioned, the compressor housing 24 includes a cylindrical inlet hole 40 having a lateral wall 42 attached to one end thereof and extending outwardly therefrom. A circumferential chamber 44 is attached to the end of the wall 42 separated from the inlet hole 40 and has a varying area around its circumference;
It increases toward the discharge from the compressor exhaust port 52.

The entrance hole 40 has a first interior wall 110 with a diameter that tapers toward the wall 42 and a step 114.
a second inner wall 110 connected to the first inner wall 110 by
It has an inner wall portion 112. This second wall 112 has a diameter that widens towards wall 42. Wall 42 has a plurality of circumferentially spaced holes 116. In addition to the compressor outlet 52, the chamber 44 is provided with a hole 118 approximately in the plane of the wall 42.

The front compressor wall insert 126 includes a cylindrical throat 128 and a circular disc 130 mounted laterally from one end of the throat 128.
The throat portion 128 includes an inner wall surface 131 with a diameter that tapers toward the disk 130 and an outer wall surface 132 with a diameter that widens toward the disk 130. A surface 132 with a widening diameter forms the entrance hole 4
The throat part 1 corresponds to the expanding surface of the inner wall part 112 of 0.
28 is inserted into and mated with the inlet hole 40. Additionally, the wall surface 131 of the throat 128 with a smaller diameter corresponds to an extension of the smaller diameter of the first inner wall 110 of the inlet hole 40 . When the insert 126 mates with the entrance hole 40, the entrance hole 4
A diameter that decreases continuously from the 0 inlet into the turbocharger is obtained.

A plurality of rivet-like projections 140 extend from disk 130 and correspond to holes 116 in wall 42. When the insert 126 is associated with the compressor housing 24 and the end of the throat 128 is associated with the step 114 of the inlet hole 40, the protrusion 140 abuts the disk 130 with the corresponding surface of the wall 42 by the hole 116.
related to. As shown in FIG.
is inserted into hole 116 and its head is modified to attach insert 126 to housing 24. Disk 130 extends beyond wall 42 to partially cover hole 118 in chamber 44. Creating a circumferential gap 146 between the outer edge of the disk 130 and the wall of the chamber 44 and defining a differential user area 148 between the centrifugal flow compressor rotor 172 and the gap 146 leading to the compressor back wall of the chamber 44. 10
0 and the disk 130.

As shown in FIG. 2, the bearing support tube 60 includes a plurality of struts 6 extending inwardly from the wall surface 64 of the inlet hole 40.
2 concentrically supported within the inlet hole 40. The turbocharger 20 has two ball bearings 162,1
The shaft 160 is rotatably supported within the bearing support tube 60 by a shaft 64. A radial turbine rotor 170 is mounted at one end of the shaft 160 and a centrifugal compressor rotor 172 is mounted intermediate the turbine rotor 170 and bearings 162,164. The shaft 160 passes through a hole 176 in the compressor back wall 100 and a labyrinth seal 178 in the turbine back wall 102.

A radial turbine rotor 170 is secured to the shaft 160 by welding or the like, and a centrifugal compressor rotor 172
is held in position on shaft 160 by retaining nut 180. Centrifugal flow compressor rotor 172 is bored to receive shaft 160 and counterbored to form hole 182 . Hole 182 has a diameter greater than the outer diameter of retaining nut 180 such that retaining nut 180 is urged into association with bottom wall 184 of hole 182 to retain the compressor rotor in position on shaft 160. . Compressor rotor shim 186 is positioned between centrifugal compressor rotor 172 and stage 188 of shaft 160 to provide precise axial positioning of centrifugal compressor rotor 172.

In the bearing support structure of the present invention, as is clear from FIGS. 2 and 3, the centrifugal compressor rotor 1
The ring 200 is fitted into the end of the bearing support cylinder 60 adjacent to the bearing support cylinder 72, and the retaining ring 202 fixed to the bearing support cylinder 60 holds the ring 20
0 does not move into the bearing support cylinder 60. An outer race 204 of the bearing device 164 is formed within the ring 200, and an inner race 206 is formed on the shaft 1.
60. Between the inner and outer races are balls 208, which form the bearing assembly 164.

The bearing device 162 is provided with an inner race 210 formed integrally with the shaft 160 and an outer race 214 formed to receive the balls 216, and a bearing support cylinder 6.
an outer ring 212 that is slidable within the outer ring 212. A compression spring 218 is installed between the outer ring 212 and a retaining ring 220 fixed within the bearing support cylinder 60.
balls 216 in the bearing devices 162, 164,
208, the outer ring 212 is biased outwardly, thus fixing the position of the shaft 160.

That is, the compression spring 218
to the left (in Fig. 3), the ball 216
is pressed against the inner race 210 formed on the shaft 160. At the same time, compression spring 218 moves outer ring 212 axially away from ring 200. Movement of ring 200 and outer ring 212 is prevented by balls 208, 216 which are prevented from moving axially past inner races 206, 210 formed on shaft 160. As a result, the balls 208 against the bearing support cylinder 60,
216 and the shaft 160 are positioned. In FIG. 3, the left end of inner race 210 (adjacent to slinger 226) and outer ring 2
The distance between the left end of the outer race 214 of No. 12 and the left end of the outer race 214 is
By making the diameter smaller than that of ball 218
This prevents the cap from jumping out toward the cap 66. Similarly, by making the distance between the left end of inner race 206 (adjacent to slinger 226) and the left end of outer race 204 smaller than the diameter of ball 208, ball 208 is prevented from popping out to the left. do.

As shown in FIGS. 2 and 3, the outer race 204 of the ring 200 faces only one side (the left) of the upper side of the ball 208. As shown in FIGS. As a result, ball 2
08 is inserted into the ring 200 from the right side, and the balls 208 and the ring 200 are brought into sliding engagement so that the balls 208 and the ring 200 can be removed from the ring 200. Therefore, the bearing device 164 is assembled by placing the balls 208 in the race 206, inserting the balls 208 into the ring 200 from the right side of the ring 200, and engaging the ring 200 with the balls 208. It will be done. Similarly, outer race 214 faces only one side (the right side) of the upper side of ball 216. Therefore, the outer ring 212 is moved to compress the spring 218, and the ball 216 is moved from the left (in FIG. 3) to the shaft 1.
Assemble the bearing assembly 162 by introducing it into the race 214 of 60. By releasing ring 212, spring 218 automatically releases ring 21.
The outer race 214 of the ring 200 is engaged with the balls 216 to form the bearing assembly 162, and at the same time the outer race 204 of the ring 200 is engaged with the balls 208 of the bearing assembly 164. Outer race 214 formed on ring 212
The gap between the inner race 210 formed on the shaft 160 is made sufficiently smaller than the diameter of the ball 216, as described above, so that the ball 216 can be moved from the position shown in FIGS. 2 and 3. This prevents it from flying out towards the cap 66.

Alternatively, by using a suitable retainer, less than an entire ball 208, 216 can be used in the bearing assembly 162, 164. Depending on the application, a self-lubricating coating can be applied to the balls,
Oil-impregnated cages, ie sacrificial cages, may also be used. In this sacrificial cage, lubrication can be achieved by consuming (sacrificing) a small amount of oil during the rotation of the balls. Mounting of the shaft 160 within the support tube 60 is completed by the connection of a cap 66 to the end of the bearing support tube 60, which closes the hole in the support tube 60 remote from the compressor rotor 172.

The use of a ball bearing arrangement with an integral inner race in this bearing system allows the use of large diameter shafts and thus provides an extremely strong shaft. Furthermore, this bearing arrangement provides an extremely tight bearing system that allows only very small axial or radial movements. As a result, such a bearing system will greatly reduce the required clearance between the compressor and turbine and the surrounding housing, and concentricity problems will be minimized.

In a preferred embodiment of the invention, bearing device 16
No. 2,164 is in a so-called oil-starved condition (a condition in which only a limited amount of oil is supplied). That is, only the amount of oil necessary for lubrication is supplied, but not enough oil is supplied for cooling. The only lubrication provided to this bearing arrangement is by wicks 222, 224 which move oil by capillary action from the sump (R) to the ramp or slinger 226. The oil supplied to the slinger 226 is forced by centrifugal force to be dumped onto the bearing arrangements 162, 164 during rotation of the shaft 160.

In this way, the use of machine oil as a lubricant for the turbocharger bearing and the associated tubing and sealing operations are eliminated. Moreover, accidents resulting from the use of dirty machine oil as a lubricant or lack of machine oil during start-up are avoided. moreover,
There is no need for oil seals, and bearing accidents caused by seal failures are also eliminated.

Alternatively, the bearing arrangement may be permanently lubricated by dense oil or grease packed within the race and around the balls of the bearing arrangement. Yet another method is to use an oil-impregnated phenol retainer to provide lubrication to the balls over a fairly long period of time. In either case, the need for wicks 222, 224 and slinger 226 is eliminated, although bearing life is reduced.

Conventional turbochargers that use journal bearings or disk-type thrust bearings require continuous lubrication of the bearings. Additionally, if the bearing is adjacent to a turbocharged turbine, the turbine is exposed to temperatures up to 1600°C, so continuous lubrication is required to ensure sufficient cooling of the bearing to prevent accidents due to overheating. There is a need. Even ball bearings require a continuous flow of oil for cooling. The ability of the present bearing system to perform well without the need for conventional large amounts of lubrication for bearing lubrication and cooling is due to the special bearing arrangement used in the present invention and to the compressor and turbine. This is a result of the relative position of the bearings.

As will be discussed below, the turbocharger with the bearing support structure can be mounted in any orientation since the bearing arrangement does not require a continuous flow of oil for bearing lubrication and cooling. On the other hand, conventional turbochargers are subject to extremely strict restrictions regarding the orientation in which they can be used.

In the turbocharger shown in FIG. 2, the journal and disc bearings of a conventional turbocharger are replaced by a highly precise globe-shaped bearing arrangement, eliminating the need for continuous lubrication flow for the bearings. Moreover, the compressor and turbine in the turbocharger shown in FIG. 2 both overhang one side of the bearing arrangement, and the turbine is placed as far away from the bearing arrangement as possible with the compressor between the bearing arrangement and the turbine. . This arrangement provides sufficient thermal insulation between the turbine and bearing equipment, so
The bearings do not heat up so much as to require conventional lubrication methods.

The arrangement of the compressor and turbine aft together in FIG. 2 not only greatly reduces the heat transferred to the bearing arrangement, but also minimizes the effects of thermal expansion, thereby reducing the need for airfoils. The gaps at the edges can be reduced. Additionally, this arrangement eliminates the need for conventional bearing housings and provides a more compact package than current straddle-mounted rotors with support bearings intermediate the compressor and turbine.

Additionally, the use of anti-friction ball bearing arrangements 162, 164 provides better control of the radial and axial movement of the compressor and turbine, which also allows for smaller end clearances. This significantly increases the efficiency of the compressor and turbine. Compared to journal bearings, anti-friction ball bearing arrangements 162, 164 also reduce the turbine work required to drive the bearings. This reduces engine backpressure, resulting in improved specific fuel consumption and enhanced rotor acceleration capability.

The present invention is directed to a nozzle area control structure 228 for selectively modifying a turbine nozzle to control the speed or pressure output of a turbocharger as shown in FIG. As seen in FIGS. 1 and 2, exhaust gas from an internal combustion engine fitted with a turbocharger is injected into the turbocharger through the turbine air intake 54 and into the turbine rear wall 102, the parallel wall 232, and the turbocharger. It leads to the blades of the turbine rotor 170 through a nozzle area 230 formed by nozzle vanes 234 . This nozzle area is controlled by a structure 228 that includes a plurality of movable nozzle vanes 234 positioned circumferentially around it and pivotable to vary the flow velocity and angle of the exhaust gas to the turbine rotor 170. . As shown in FIGS. 2 and 4, vane 234 includes trunnions 236 and 238 extending from opposite sides thereof. Trunnion 236 passes through turbine rear wall 102 and attaches to operating lever 240 . A trunnion 238 extends into wall 232.

A nipple 242 is provided at one end of the operating lever 240.
form. These nipples are connected to control ring 2
46 into a radial slot 244 formed therein. Control ring 246 and operating lever 240
The compressor rotor 172 and the turbine roller 17
The air gap gap 247 is located midway between zero and zero. Control ring 246 is positioned concentrically about the axis of shaft 160 and is received on a cylindrical surface 248 extending from compressor back wall 100 .

In FIG. 2, the control ring 246 is shown as an inner ring 25 formed with an inner race and an outer race and receiving a plurality of balls 254 therebetween.
0 and an outer ring 252. inner ring 2
50 is a cylindrical surface 24 extending from the compressor rear wall 100
8, the outer ring 252 rotates angularly relative to the inner ring. As shown in FIG. 4, the rotation of the outer ring 252 causes each operating lever 240 to rotate about the axis of the trunnions 236, 238, thereby causing each nozzle vane 234 to rotate simultaneously. As shown in FIGS. 4 and 5, one of the operating levers 240 is connected to the extension 262.
It is equipped with Control rod 86 includes a threaded eyebolt 86a attached to extension 262 by an axle pin. The opposite end of eyebolt 86a is threadedly received within control rod 86 and is adjustable therein to permit readjustment of the nozzle vane about its axis of rotation. Movement of the control rod 86 pivots the operating lever 240 so that the control ring 2
46 is rotated, thereby rotating each of the other operating levers 240 and the nozzle blades 234 attached thereto.

FIG. 5 shows an actuator 80 used in an Otto cycle engine. In this system, the control rod 86 is a piston-type actuator 8
Controlled by 0. Alternatively, these pistons may be replaced by diaphragms. Actuator 80
is controlled by compressor discharge pressure sent to controller 84 through tubes 88A, 88B. The increased pressure entering the actuator 80 causes an extension of the control rod 86 and a corresponding opening of the nozzle area of the compressor.

As shown in FIG. 5a, the control rod 86 is connected to the controller 8.
4 to change the position of the nozzle blade 234. The controller 84 has a housing 264 forming a large cylinder 265 and a small cylinder 266.
It is equipped with A large piston 268 with a front surface 268a and a rear surface 268b is placed within the cylinder 265. A small piston 270 with a front surface 270a and a rear surface 270b is placed within the cylinder 266. Attach the front plate 272 to the housing 2 with appropriate bolts.
64 and includes a hole 273 for receiving the control rod 86 into the housing. The piston 270 is the control rod 8
6, while the piston 268
is slidable relative to the control rod 86, and the control rod 86 adjacent to the front surface 268a of the piston 268
It is held in place only by a ring 274 which is fixedly attached to.

A spring 276 is provided around the control rod 86 and placed between the piston 268 and the front plate 272. A larger spring 278 is placed around control rod 86 and between piston 270 and piston 268. The air pipes 88B and 88A are connected to the rear surface 2 of the piston 270.
70b, and the piston 2
The cylinders 265 are connected to the front surfaces 268a of the cylinders 68, respectively. Air tubes 88A, 88B have one end attached to the engine's intake manifold and the other end attached to controller 84 to direct manifold pressure to front 268a and rear 270b of pistons 268, 270, respectively. Air pipe 86
are the rear surfaces 26 of the pistons 268 and 270, respectively.
8b and the front surface 270a are supplied with atmospheric pressure.

An annular groove 280 and a bore 282 through a guide cylinder 284 projecting from the rear wall 268b of the piston 268 ensure transmission of atmospheric pressure along the entire rear surface of the piston 268. Similarly,
Radial groove 28 on the rear surface of the piston 268
6 facilitates the transmission of pressure to the entire rear face of the piston.

Controller 84 controls nozzle vanes 23 to turbine roller 170 at both low and high manifold pressures.
The plan is to extend control rod 86 to open 4 and retract control rod 86 to close nozzle vanes at intermediate manifold pressures. More specifically, while operating in the closed throttle position when the suction manifold is at a vacuum, such as 5 to 8 psi absolute pressure, the controller opens the nozzle vanes 234 so that the engine has as little back pressure as possible. It acts to extend the control rod 86 so as to work at the same time.
This results from the high atmospheric pressure (approximately 14.7 psi absolute) from tube 89 acting on the rear face 268b of piston 268 compared to the pressure from tube 88A, which is significantly lower than atmospheric pressure.

While the net pressure difference across piston 270 creates a force that allows control rod 86 to be retracted, the surface area of pistons 268, 270
It is planned that the outward force exerted on piston 270 is sufficient to overcome both the retraction pressure on piston 270 and the spring force of spring 276.

The spring rate of spring 276 and piston dimensions are:
When the manifold pressure approaches atmospheric pressure, such as 12 psi absolute, control rod 86 is retracted into controller 84 to close the nozzle vanes, thereby reducing the speed of turbocharger 20 and the pressure available from the turbocharger compressor. It is planned to increase the ratio. The controller 84 schedules the closing of the vanes 234 to begin just before atmospheric pressure is reached, since this suction manifold pressure corresponds to an open carburetor throttle and therefore eliminates the need for turbocharger boost. This is because it shows that

As shown in FIG. 5a, as the manifold pressure in the tube 88A acting on the front surface 268a of the piston 268 approaches atmospheric pressure, this pressure, together with the action of the spring 276, is applied from the tube 89 to the rear surface 268b of the piston 286. Control rod 86 overcomes atmospheric pressure
into the controller 84. The pressure differential across piston 270 still exerts a net force that retracts piston 270 into its cylinder, thus assisting in retraction of control rod 86 into controller 84 .

As the manifold pressure increases above atmospheric pressure, such as 20 psi absolute, tube 88B
The manifold pressure acting on the rear surface 270b of the piston 270 through overcomes the spring force of the spring 278, causing the control rod 86 to move outwardly from the controller 84.
The vanes 234 are opened again, thereby preventing excessive pressurization. Control rod 86 extends from controller 84 under this pressure, but piston 268 is held in the fully retracted position within cylinder 265 because piston 270 is under the force of spring 278. This is because control rod 86 has the ability to move freely relative to piston 268 when moving toward piston 268 .

The controller is therefore configured to open the nozzle vane 234 at a manifold pressure having an absolute reading less than or equal to a predetermined value below atmospheric pressure, as determined according to the schedule of the controller and components therein. system. Similarly, the controller closes vanes 234 at manifold pressures ranging from some predetermined value below atmospheric pressure to some value above atmospheric pressure, and closes vanes 234 at manifold pressures above a predetermined value above atmospheric pressure. open. Of course, the particular pressure at which the vanes open and close can be easily varied by varying the area of the pistons 268, 270 and by varying the spring rate and initial deflection of the springs 276, 278.

For diesel cycle engines, a solenoid can be used to control the piston 268, in which case the solenoid keeps the shaft 86 extended during light load operation of the engine and when heavy loads are required. If this happens, release the shaft. Upon releasing the shaft, spring 276 retracts shaft 86, closing the nozzle and increasing turbocharger speed. When sufficient pressure is delivered from the suction manifold to face 270b of piston 270, shaft 86 extends to open the nozzle and thereby control manifold pressure, as described above.

When operating variable area turbine nozzles,
A control signal, such as compressor discharge pressure, is communicated to actuator 80, which extends or retracts control rod 86 as appropriate in accordance with the signal to the actuator. This allows rotation of the actuating lever 240 attached to the control rod 86 and control ring 2.
46 simultaneous angular rotations of the outer ring 252 occur. Further, the rotation of the outer ring 252 rotates each actuating lever 240 and sets the angle of the nozzle blade 234 accordingly.

FIG. 4 shows the nozzle vanes 234 in the closed and open positions using dashed and solid lines. It can be seen that each nozzle vane is set through movement of a single control rod 86 controlled by a single actuation lever 240 by rotation of control ring 246.

Traditional uncontrolled turbocharging typically produces a boost pressure that is directly related to compressor speed;
This compressor speed normally corresponds to engine speed.
Therefore, the boost pressure in such conventional turbochargers is lower when the engine speed is slow than when it is fast. Therefore, at low engine speeds, conventional turbocharging is not effective in improving engine performance. Additionally, during acceleration, there is a significant "lag time" before the turbocharger reaches speed to produce sufficient boost pressure to effectively improve engine performance.

The turbocharger shown in FIG. 2 can be controlled to provide boost pressure at low engine speeds. As a result, engine performance can be improved at lower engine speeds and the lag time associated with conventional turbocharging can be eliminated or reduced.

This result is achieved through variable turbine nozzle control that maintains the turbocharge turbine and compressor speeds at a level that provides the required boost at all engine speeds. Sixth,
As seen in FIG. 7, actuator 80 is replaced by control unit 400. The control unit 400 includes a magnetic sensor 402 coupled to a monitor 403 via a pulse-to-voltage converter 404. Rotation of shaft 160 is monitored by sensor 402 in conjunction with disk 160'' mounted on shaft 160. Disk 160'' has a plurality of discontinuities. The speed of shaft 160 is determined, in a well-known manner, by converting the electrical pulses produced by the movement of the disk relative to magnetic sensor 402 and then inputting these pulses to pulse-to-voltage converter 404. It is determined. This voltage, which is proportional to the rotational speed of the disk 160'' and shaft 160, is converted to shaft speed by a monitor 403. The monitor 403 is coupled by suitable leads to a proportional controller 414.
is a power supply 416 and a nozzle area control unit 415.
Connect to. This unit 415 has a control rod 86
and rotate the lever extension 262 to control the movement of the blades 23 in the inlet area of the turbine rotor 170.
Change the direction of 4.

Controller 414 is a microprocessor,
It includes a conversion system that receives analog input from a sensor 417 and a speed monitor 403 for conversion to a digital format that can be used by the microprocessor. This microprocessor may be of the general type currently used in automobile ignition systems. Alternatively, the microprocessor that can be incorporated into this system may be the microprocessor model 8080 manufactured by Intel. Proportional controller 414 includes an amplification system for amplifying the signal provided by the microprocessor for controlling control unit 415.

This microprocessor determines the turbine speed and therefore the boost pressure produced by the turbocharger compressor as a function of several parameters, such as engine speed, throttle position, manifold suction pressure, engine torque, ambient temperature, or transmission gearing. , you can configure its programming to control. As shown in FIG. 6, sensor 417 measures each of these engine parameters and sends corresponding analog readings to proportional controller 414, where the parameters are digitally converted for use by a microprocessor. converted to form.
The microprocessor can, of course, be programmed to position the turbine inlet nozzle to control compressor speed according to any direct or variable relationship of parameters.

Proportional controller 41 shown in FIG. 6 by FIG. 6A
4 will be explained in detail. Multiple sensors 417a-g
measures parameters related to the operation of an automobile engine. Each of the sensors 417 generates an analog signal along with the turbine speed monitor 403, which is connected to a respective analog-to-digital converter 424a-h.
can be conveyed to. The signals generated by each of these sensors are converted in a corresponding analog-to-digital converter into a series of digital words representative of the numerical value of the parameter being measured.

The digital signal produced by analog-to-digital converter 424 is selected by a microprocessor contained within proportional controller 414. Microprocessor 426 receives digitized signals representative of the parameters being measured by sensor 417. This microprocessor evaluates the combination of signals received and then
A signal corresponding to the best control position for 34 is generated. The output signal produced by microprocessor 426 is sent to digital-to-analog converter 428.
and this transducer 428 generates a corresponding nozzle control signal. The output of digital-to-analog converter 428 is connected to a driver circuit 429 that amplifies the nozzle control signal. The amplified signal is communicated to nozzle area control unit 415 which selectively positions nozzle vanes 234. Driver circuit 429 not only provides amplification for the nozzle control signal, but also isolates proportional controller 414 from spurious signals generated in the area of control unit 415 and proportional controller 414.

Proportional controller 414 and sensor 417 utilize techniques previously developed for automotive engine control for spark timing and fuel metering. Existing systems of this type are described in the book Automotive Electronics Gets The Green by Gerald M. Walker.
Light” ELECTRONICS, Volume 50, No. 20
(September 29, 1977), pages 83-92.
In this way, the engine designer can obtain the required boost pressure relative to one or more of these or other parameters. Of course, the parameters listed above are only indicative of some of the parameters that a designer may wish to use to select the best turbocharger compressor speed throughout the full range of engine and turbocharger motion. .

The simplest arrangement is represented by simply keeping the turbine speed and therefore the compressor speed at a constant level throughout the entire operation of the engine. In this condition, a setpoint speed is selected for controller 414. Controller 414 sends a signal to actuator 418 in response to the comparison of the set point speed to the speed indicated by monitor 403. Monitor 403
If the axis speed monitored by actuator 4 is less than or equal to the set point value of controller 414, controller
18 to retract rod 86, resulting in rotation of vane 234 and closure of the turbine nozzle. When the turbine nozzle is closed, the turbine speed increases and the rotational speed of shaft 160 increases. If the shaft speed indicated by monitor 403 is greater than the set point value, a signal from controller 414 to actuator 418 causes extension of rod 86 and corresponding opening of the turbine nozzle to cause the turbine and its This will reduce the speed of the shaft 160 attached to. By continuing such operations, the speed of the turbine and compressor is controlled by the controller 41.
4 can be controlled according to a predetermined set point value programmed into the set point value. Thus, regardless of engine speed and throttle position, by varying the position of vanes 234 a constant compressor and turbine speed can be maintained and the flow rate of exhaust gas hitting turbine rotor 170 can be controlled.

If you use the turbocharger shown in Figure 2,
A relatively high compressor speed can be maintained regardless of engine speed. This is possible due to the possibility of changing the area of the turbine nozzle. Therefore, no boost pressure lag time is observed when accelerating the engine, since there is no need to increase the compressor speed from a relatively low speed to a high value that is continuously maintained by the turbocharger. It is. Furthermore, by maintaining high boost pressure at relatively low engine speeds, the turbocharger provides improved engine performance at both low and high engine speeds.

The turbocharger operation just described assumes that the turbocharger compressor can be used over a wide flow range. Although it may offer some advantages over conventional turbocharging as discussed below, the benefits provided by the turbocharging variable area turbine nozzle shown in FIG. This can be significantly expanded by using a compressor that can. That is,
Conventional turbochargers, which have suffered from stall lines that prohibit compressor operation over a wide flow range, do not benefit as much from the use of variable area turbine nozzles as in the turbocharger shown in Figure 2. Will.

Accordingly, the turbocharger shown in FIG. It is. Furthermore, the control ring and actuating lever are located in unused air goat space between the turbocharger turbine and the compressor, forming a very compact unit. Furthermore, the special location of the linkage insulates the bearing arrangement from the heat experienced by the turbine and its surrounding area. In this way, the bearing arrangement can be made oil-free or with limited lubrication. This also obviates the need for conventional lubrication through the use of machine oil and the associated tubing required for such lubrication methods. Needless to say, this makes the turbocharger economical, inexpensive to manufacture and operate, and reliable.

Furthermore, the use of ball bearing supports and the corresponding elimination of thrust bearings and disk-type bearings provides a more controlled or stable turbine-compressor rotating system, which provides both turbine and compressor support. Only narrow gaps are allowed between these and surrounding structures. For example, the height of turbocharger compressor blades used in relatively small internal combustion engines is on the order of 0.2 inches. Using the ball bearing arrangement described above, the compressor construction requires only 0.005 inch clearance, or only 2.5% of the total blade height. On the other hand, if journal bearings or disk-type bearings are used, significantly larger gaps (usually about
0.015 inch) would be required. As a result,
The turbocharger shown in FIG. 2 is particularly adapted to turbocharger construction for small internal combustion engines where low blade heights require narrow gaps between the blades and the surrounding structure.

The use of ball bearings and the associated lower friction losses compared to journal or disc type bearings allows turbochargers to generate less turbine horsepower without losing a significant portion of the turbine horsepower to bearing losses and friction. Enabling effective use for the first time. Considering that journal bearings and disk type bearings account for 4 to 5 horsepower losses in friction, the friction losses with the arrangement of the present invention will be on the order of 0.1 to 0.4 horsepower.

FIG. 8 shows compressor rotor 172, retaining nut 180 and compressor-turbine shaft 160 spaced from turbine rotor 170. turbine rotor 1
70 is attached to one end of shaft 160 by welding or other suitable permanent attachment means. The shaft 160 has an enlarged bearing surface 160 before being attached to the rotor 170.
a and a stepped portion 188 that reaches the small shaft diameter portion 160b. As described above, the races 206 and 210 are formed directly on the shaft 160.

During assembly, the shaft 160 is inserted through holes in the compressor back wall 100 and the turbine cladding 102. Shim 186 is positioned on shaft 160 in association with stepped portion 188 of shaft 160. Compressor rotor 172 is associated with shaft 160 and positioned on shaft portion 160b. The retaining nut 180 is then attached to the shaft portion 1
60b is pushed into the hole 182 of the rotor 172 and associated with the bottom wall 184 of the hole 182. Nut 180 consists of a sleeve with a smooth inner bore 180a. A hole extending through nut 180 forms an interference fit with portion 160b of shaft 160. This interference fit is, for example, on the order of 0.001 inch.

Since the inner race is formed directly on the compressor-turbine shaft, the shaft must be heat treated to extremely high hardness. As a result, the retaining sleeve secures the compressor rotor 172 to the shaft 160 without the need for grinding or cutting threads into the hardened shaft. In this manner, the cost and problems associated with forming threads on heat treated shafts are eliminated. Additionally, because the shaft is substantially hardened, the retaining ring 180 can be pushed against the shaft and pulled out without damaging the shaft surface.

As shown in FIG. 8, nut 180 is threaded about its outer surface. The hole 182 has a gap 420 (9th
The diameter is large enough to accommodate the This gap 420 allows the insertion of a suitable internally threaded tool to pull the nut from the shaft for removal of the compressor rotor.

A spring system is required between the retaining nut and the compressor rotor to maintain an axial force on the compressor rotor as the turbocharger components expand or contract. Unlike the case of internally threaded nuts associated with threaded shafts, the retaining nuts of the present invention do not support high compressive loads within the compressor rotor or tension within the compressor-turbine shaft when installed in place. It does not have the ability to cause Accordingly, in FIG. 10, a conical or Bellevueille spring 422 is inserted between retaining nut 180 and compressor rotor 172.

The components in FIG. 10 are shown in FIGS. 2 and 8,
Since these parts are the same as or equivalent to those shown in Figure 9, the same numbers marked with (') will be used to indicate parts that are equivalent to or similar to those of the embodiments shown in the latter figures. . In FIG. 10, axis 16
0' passes through compressor rotor 172'. A retaining nut 180' holds the shaft 160' at the end of the shaft 160'.
Bellevue spring 422 attached to and nut 18
0' and the wall 184' of the bore 182' of the rotor 172'. Bellevue spring 422 is originally in a compressed state when attaching retaining nut 180' to shaft 160'. The connection of nut 180' to shaft 160' is sufficient to overcome any expansion forces created by Bellevue spring 422 between nut 180' and rotor 172'. Instead, the compressive force of spring 422 between rotor 172' and nut 180'
2' induces an axial load. Thus, with spring 422 in its normal position, contraction or expansion within shaft 160' or rotor 172' will cause nut 18
The coupling force between rotor 172' and rotor 172' does not become zero.

FIG. 11 shows the compressor housing 24, the compressor front wall inserter 126, and the respective molds used to die cast each of these two pieces. Conventionally, the compressor housing of many turbochargers is made by sand casting so that a circumferential chamber into which the compressed air is directed into the intake manifold of the engine with which the turbocharger is used is provided. It may be formed in a generally enclosed configuration with a narrow circumferential gap for receiving compressed air to allow for the formation of a diffuser. Making the compressor housing in one piece by sand casting is more expensive than making it by die casting and leads to the creation of a less precise structure with a rough wall structure within the circumferential chamber. However, in the past when turbocharger compressor housings were made by die casting, it was not possible to make the front diffuser wall as was possible in sand casting.
The reason for this is that the die casting mold had to be equipped with appropriate inlets and outlets.

As shown in FIG. 11, a two-component die-cast compressor housing can be provided that has advantages in assembly that are conventionally available only by sand casting. The compressor housing 24 has an outer mold 290,
Inner mold 292, core mold 294 and cap mold 296
Form using. Inner mold 292 is given a raised profile to define circumferential chamber 44 . Also, this inner mold 292 has a protruding portion 300.
The inlet hole 40, the vane 62, and the bearing support cylinder 60 are formed in combination with the corresponding protruding portions from the outer mold 290. Core mold 29
4 and the cap mold 296 are the exhaust port 52 of the compressor.
used to form.

As is clear from FIG. 11, molds 290, 29
2,294,296 cooperate with each other to enable die casting of the compressor housing 24.
The molds 290, 292 are provided with abutment surfaces to form a separation line 310 in the housing 24. Type 320, 32
2 cooperate to create the compressor front wall insert 126. The molds 320, 322 have abutting surfaces that interlock to form a separation line 324 on the disk 13.
Create an insert 126 that has an outer edge of 0. As best seen in FIG. 11, the throat 128 is provided with a notch 326 which allows the insert 1 to
26 to receive struts 62 when installed within housing 24.

Figure 12 and cutaway diagrams 13, 14, and 15 are
The positioning and shape of the blade 62 is shown. 1st
3-15, each strut 62 has a leading edge 330 and a trailing edge 332 with a thicker intermediate center cut 334. In FIGS. In each case, the thickest central cut is as indicated by line 336 and
Molds 290, 29 used to form housing 24
It forms a line of separation between the two. That is, the pillar 6
2 is formed by die casting using molds 290, 292 to produce the required air-oil shape of the leading and trailing edges separated by a thicker central cut between them. This shape is the 13th,
As shown in Figures 14 and 15, the shape of the present invention greatly facilitates the entry of air into the compressor inlet area and can be cast using known die casting techniques.

To ensure that the airfoil shape of the strut 62 tapers from the thicker center cut to the thinner leading and trailing edges of the vane, the die-casting mold has inlet holes 40 from both the leading and trailing ends. must be inserted inside. That is, as shown in FIG. 11, the second inner wall 112 (FIG. 2) is formed by the protrusion 300 of the mold 292, while the first inner wall 110 (FIG. 2) is formed by the mold 290. It is formed as a result. Both inner walls 110, 112 each widen outwardly to allow the die casting mold to be removed after the piece is formed.

In this way, the formation of the required shape of the support column 62 is as follows:
Requires a diverging diameter wall 112 in the entry hole 40 for removal of the die casting mold. However, it is important that the inlet hole 40 has a diameter that decreases continuously from the inlet toward the compressor rotor 172. This is done by using wall inserts 126. The wall insert 126 is also die-cast and extends from the inner wall 112 of the entrance hole 40.
and interlocking outer surfaces 132 (FIG. 2) having divergent diameters corresponding to the divergent surfaces of the .
The inner wall surface 131 of the throat 128 is formed with a diameter that tapers toward the disk 130 in response to the extension of the tapered diameter of the first inner wall 110 of the inlet hole 40 . That is, the wall insert 12
6 in combination with the inlet hole 40 provides a diameter that tapers continuously inwardly from the inlet of the hole 40 toward the compressor rotor 172 (FIG. 2).

In this way, a two-piece construction forming the compressor housing for the turbocharger is obtained, which can be made together by die-casting. These two components form a compressor housing with a plurality of struts 62 that support a bearing support cylinder 60 during assembly, with thicker struts that decrease toward a thinner leading edge and a thinner trailing edge. It has a vane with a mid-cut. Furthermore, the housing defines an inlet nozzle that decreases continuously from the inlet of the inlet hole toward the compressor rotor. Additionally, the compressor housing defines a front diffuser wall that completes a circumferential chamber through which compressed gas is communicated to the compressor exhaust.

In FIG. 16, different shapes of wall inserts 126 are shown to be interchangeable for use with corresponding compressor rotors. Wall insert 126a provides more restricted airflow within the turbocharger;
Wall insert 126b forms a larger compressor rotor and provides a larger air flow to the turbocharger. It will be appreciated that such structural changes can be made simply by attaching different wall inserts 126 to the standard compressor housing 24. That is, a variety of different turbochargers, each with different flow performance, can use a standard compressor housing 24, each with one of several possible shapes for the wall insert 126 with a corresponding compressor wheel. is created by selecting. This feature is significant in that the wall insert 126 is a simple component of the compressor housing.

In conventional turbochargers, whether die-cast or sand-cast, retrofitting involves making an entirely new casting for the new structure. However, as previously mentioned, compressor designs with different air flow to air pressure ratios can be simply substituted by placing the wall inserts of the different designs in cooperation with a new or modified compressor wheel. can be achieved.

It is therefore possible to provide a turbocharger compressor housing that can be die cast as described above. The compressor rotor housing includes a compressor housing and a front wall insert associated with the housing. Such a two-part construction allows each of these two parts to be die-cast using standard die-casting techniques. A compressor housing is the combination of these parts with a circumferential chamber closed on all sides except for a circumferential gap for receiving compressed gas from the compressor rotor. It is made as. A circumferential passage is formed in the gap leading to this chamber, in which the gas accelerated is diffused and increases the static pressure.

Additionally, the compressor bearing support tube is cast concentrically with the inlet hole and supported therein by a plurality of vanes extending from the wall of the inlet hole. The vanes are formed with leading and trailing edges separated by a thicker center cut. This is done by using a mold with a separation line in the generally thicker cross-cut area to allow die-casting of the compressor housing.

Die-casting the compressor housing in two components is preferred for turbochargers in which the compressor rotor and turbine rotor project to and are supported on one side of a shaft which rotates on a ball bearing arrangement. Although the present invention has been described and applied, it will be obvious to those concerned that it can be easily applied to a turbocharger having a conventional structure in which a compressor rotor and a turbine rotor are supported at opposite ends of a bearing support device. . For this special purpose, the compressor housing as described above can be used as is by simply modifying the bearing support tube.

An important feature of the improved turbocharger with the particular design configuration of the air inlet 50 obtained by selecting the shapes of the struts 62 and the bearing support tube 60 is that it can be used over an extremely wide range of compressor speeds and flows. This is the ability to operate without experiencing the surging lines normally found in conventional turbochargers.
This extremely significant and important feature is
This will be explained using a performance graph known as a compressor map shown in Figure 8. FIG. 17 is a typical compressor map for a conventional turbocharger, and FIG. 18 is a compressor map for the improved turbocharger. These compressor maps show the relationship between cubic feet per minute of airflow versus the ratio of discharge pressure to inlet pressure for a turbocharger compressor.

FIG. 17 shows the constant speed line 430 of the compressor.
and isentropic efficiency line 432 are shown. Surging line 434 defines the operating point, to the left of which the turbocharger compressor cannot operate to produce a uniform air output. Although the turbocharger components can be modified to provide stable operation at lower turbocharge speeds as the surging line moves toward the vertical axis, i.e., at lower engine speeds, this Such movements reduce the unit's high velocity capabilities. Accordingly, turbochargers have traditionally been downgraded to provide increased performance only at higher engine speeds.

Thus, it is possible to provide a turbocharger capable of producing uniform power at extremely low turbocharged airflows, while eliminating the surging line limitations encountered with conventional turbochargers. As is apparent from FIG. 18, the compressor performance map for the improved turbocharger has a compressor constant speed line 440 and an efficiency line 442.

As can be seen from the compressor map of FIG. 18, the improved turbocharger compressor continues to operate without irregularities in compressor output down to extremely low airflow values. In fact, the improved turbocharger does not experience surging lines comparable to those shown in the compressor map of FIG.

17 and 18, the width of the operating range of the conventional turbocharger can be compared with the improved turbocharger. According to FIG. 17, at a compressor pressure ratio of 1.9, the airflow range varies from a high of 350 cubic feet per minute at 60% efficiency to as low as 200 cubic feet per minute at the surging line. Therefore, the ratio of high flow rate to low flow rate (350÷200) is
It becomes 1.75. According to FIG. 18, at a pressure ratio of 1.9, the improved turbocharger has an efficiency of
It operates from a flow rate of 275 cubic feet at 60% to 60 cubic feet per minute without encountering surging lines. The ratio of the high and low flow rates (275÷60) is 4.58, which is more than 2.5 times the increase in conventional turbocharger.

These results are significant in that turbocharging can be used effectively to improve performance at both low and high engine speeds. Such advances, when combined with the aforementioned features of providing a constant compressor speed for varying engine speeds or a constant boost at varying engine speeds, provide even higher desired pressure ratios at lower engine speeds. It is also of greater significance in that it is obtained without the surging or turbulence problems normally experienced with lower compressor flows in conventional turbochargers.

These significant advances provided by the improved turbocharger are primarily attributable to the compressor air inlet geometry in conjunction with advanced radial compressor technology, such as back-curved blade arrangements. It is something that should be done. As already mentioned, the bearing support tube 60 includes a plurality of struts 6 within the cylindrical entrance hole 40 extending from the inner wall of the hole to the bearing support tube 60.
It is supported by 2. In one embodiment, three struts extend from the inner wall of the inlet hole to a bearing support tube that is located concentrically within the air inlet of the compressor. As a result, the inlet is divided into one or more inlet channels. Furthermore, these channels
Proportionally long in the flow direction compared to the width dimension. Each strut also has a profile with leading and trailing edges that are narrower than the intermediate region of the strut. As also shown in FIG. 2, the interior wall of the inlet hole 40 has a diameter that tapers from the inlet end to the compressor rotor 172.

This structure stabilizes the backflow conditions normally encountered at low compressor flows. By stabilizing the backflow conditions, the compressor can continue to operate to produce a steady air flow at progressively lower compressor flows than previously possible with conventional turbochargers.

FIG. 19 shows a vertical sectional view of a modification of the turbocharger illustrated in FIGS. 1 to 16. 19th
The turbocharger shown in the figure is the axial flow stage compressor rotor 4.
50 was installed inside the compressor inlet.
This is the same as the turbocharger illustrated in FIGS. Since many parts of the variant shown in FIG. 19 and the turbocharger illustrated in FIGS. 1-16 are substantially the same, the same reference numerals have been used for similar or corresponding parts.

In FIG. 19, the compressor inlet is defined as an inner wall 454 that partitions an inlet passage 456 to the centrifugal compressor rotor 172.
It has been modified to include a larger diameter inlet 452 with a larger diameter. The plurality of stators 458 are connected to the inlet 4
52 and supports a bearing support 460. Axial stage compressor rotor 450 includes centrifugal compressor rotor 17 in each bearing assembly 162, 164.
It is attached to the end of shaft 160 on the opposite side of 2 and held by a suitable nut 462. The compressor rotor 450 has a boss 464 and a support 4.
68 and a plurality of support columns 466 attached between the two. Boss 464 is associated with step 470 of shaft 160 and is located between step 470 and nut 462 .

As is clear from FIG. 19, the air entering the inlet 452 is compressed by the axial stage compressor, and the air enters the flow path 452.
and is conveyed to the centrifugal compressor rotor 172. 2
The use of a stage compressor is important in that the use of a two stage compressor provides a wider operating range compared to that available from a stage compressor. For example, when using a two-stage compressor, the first stage is used to obtain an inlet pressure to discharge pressure ratio of 1.5 to 1.6, while the second stage is used to obtain a pressure ratio of 3.0 to 3.5. These combined pressure ratios effectively provide pressure ratios on the order of 4.5 to 5.6. Although such pressure ratios can be obtained with a single stage, the higher compressor speeds required to produce such pressure ratios require the use of higher quality and more expensive compressor and turbine components. Moreover, the flow range is severely restricted. A two-stage unit provides higher pressure ratios without this additional expense or drawback. Additionally, by mounting the axial stage compressor on the opposite side of the bearing support from the centrifugal compressor and turbine rotors, the system is balanced and certain adverse effects are avoided.

20, 21, 22 and 23 show a new method of attaching the turbocharger illustrated in FIGS. 1 to 16 to an internal combustion engine. Second
0 and 21 show a hard sectional view and a plan view, respectively, of the turbocharger 20 attached to the V-6 engine 444. Of course, this same application can be applied to V-type engines with any number of cylinders. Engine 444 represents a conventional V-engine with a right cylinder group 445 and a left cylinder group 446. The cross section shown in FIG. 20 is a cross section passing through the exhaust valve 448 of the right cylinder group 445 and the intake valve 449 of the left cylinder group 446. The turbocharger 20 is mounted in a rigid orientation with the rotational axes of the turbine and compressor in a rigid orientation, which is not possible with conventional turbochargers. turbo charger 2
0, oil is delivered to the bearings by means of a wick and a corresponding structure that does not require a continuous flow of lubricant to the bearings, so that the axes of rotation of the compressor and turbine are aligned vertically or between vertical and horizontal lines. The turbocharger can be oriented at any angle. On the other hand, conventional turbochargers, which normally require a flow of lubricant to and from the bearings, have not been successfully operated in a vertical orientation, as exemplified by the configurations of Figures 20 and 21. do not have.

As shown in FIGS. 20 and 21, the engine exhaust gas is guided to the turbocharger 20 by a manifold 451. In this case, the exhaust is provided by each individual exhaust valve 448.
From the turbine rotor 17 of the turbocharger 20
Leads directly to 0. Turbocharger compressor exhaust is routed through each cylinder of engine 444 by pipe 453 through manifold 447 to intake valve 449.

In operation, air is drawn into the turbocharger 20 at the inlet 50 and discharged by the compressor rotor 172 through the compression stub 453 into the manifold 447 and into each cylinder of the left and right groups of the engine 444. Exhaust gas from engine 444 is delivered by manifold 451 to turbine rotor 170 at spaced locations around it and then discharged through exhaust pipe 455 . This interval allows
The high pulsating energy created when exhaust valve 448 first opens can be utilized by the turbine.

As is apparent from FIGS. 20 and 21, the turbine rotor 170 is comprised within a turbine housing having a circumferential chamber having variable nozzles therein substantially aligned with a plane orthogonal to the axis of rotation of the rotor. It is installed on. Furthermore, each engine exhaust is aligned with a common horizontally oriented plane. As shown, in place of the conventional turbine housing commonly used, a manifold 451 is used which acts both as an exhaust manifold and as a turbine housing. As mentioned above, since the turbocharger can be used in a vertical orientation, each exhaust port can be positioned in the plane of the circumferential chamber of the turbine housing by locating the turbine shaft axis approximately perpendicular to the plane of each exhaust port. planes can be aligned substantially parallel. In this way, the exhaust air from the exhaust port is more directly and effectively injected into the turbine housing to drive the turbine rotor.

Conventional turbochargers, on the other hand, require the use of extremely complex manifolds to direct exhaust gas from the exhaust port to the turbine rotor. The turbocharger 20 further includes an exhaust manifold that directs the exhaust gas from each exhaust port from the engine to approximately equally spaced locations around the turbine rotor and directs the exhaust gas to different locations around the periphery of the turbine rotor. send. Because the turbocharger 20 can be vertically oriented, the turbocharger 20 can also be located close to the engine between each group of cylinders. Furthermore, the exhaust manifold directs the exhaust air directly from each cylinder to the turbocharger turbine with very little loss of energy before applying the exhaust to drive the turbine.

The carburetor is of course mounted directly above the turbocharger and delivers a fuel-air mixture into the inlet 50. Similarly, the turbocharger can be used in fuel injection engines or diesel engines that deliver fuel directly to each engine cylinder.

22 and 23 show the use of a vertically mounted turbocharger 20 in a parallel four cylinder engine. turbo charger 2
Of course, 0 can also be used in parallel engines with more or less than 4 cylinders. 22 and 23, the turbocharger 20 includes an exhaust manifold 482 that communicates with the turbocharger turbine between the exhaust valves of each engine cylinder.
It is connected to engine 480 by. Air is drawn into the turbocharger 20 at the inlet 50, compressed by the turbocharger compressor rotor 172, and discharged through the manifold 484 into the intake valve of each engine cylinder. As seen in FIG. 23, the exhaust manifold 482 is configured to direct the gases that drive the turbocharged turbine to the turbine at spaced locations around its periphery.
Additionally, manifold 482 is oriented to facilitate rotation of the turbine clockwise in FIG. 23. Similarly, as can be seen in FIG. 21, a similar manifold is provided when the present turbocharger is mounted intermediate the two cylinder groups of a V-engine.

Thus, a turbocharger is obtained which can be oriented with the rotation axes of the turbine and compressor vertically or at any position intermediate between horizontal and vertical. This is primarily possible because the turbocharger 20 lubricates the bearings around which the compressor and turbine shafts rotate. Because the turbocharger 20 can be oriented in this manner, it is ideally suited for use in conventional engines of all types and constructions.

The turbocharger described in connection with FIGS. 1-16 can be easily converted into an effective turbojet or turboblower engine as shown in FIGS. 24, 25, 26, 27, and 28. can. In these embodiments, the vertical cross section of only the upper half is illustrated, but it goes without saying that the lower half is almost the same as the upper half shown.

Turbojet 500 is illustrated in FIG. The turbojet 500 includes a tubular inlet nozzle 504, a main housing 506, and a combustion chamber rear wall plate 50.
The external housing 502 is comprised of 8. The rear wall plate 508 includes integral reinforcements 510 and is attached to the main housing 506 by bolts 512. Turbine exhaust nozzle 514 is supported within main housing 506 by being attached to the end of rear wall plate 508 opposite main housing 506 by bolts 515 .

The bearing support cylinder 516 is connected to the tubular inflow nozzle 50
A plurality of pillars 51 extending from the inner wall surface 520 of No. 4
8 within the inlet nozzle 504. This structure is similar to that illustrated in FIGS. 1 to 16 with respect to the bearing support cylinder 60 and its attachment within the inlet 40 of the turbocharger 20. Cap 522 is attached to the end of bearing support cylinder 516. Note that, as shown in FIG. 24, the compressor rear wall 524 and the turbine rear wall 526 are located within the main housing 506 and fitted with appropriate threaded pieces 527.
are attached to each other. In the embodiment illustrated in FIG. 24, a fixed stationary vane 528 is attached between the main housing 506 and the compressor rear wall 524 by a suitable threaded piece 529. combustion liner 530
is the turbine rear wall 526 and the turbine exhaust nozzle 51
It is attached between the horizontally extending portion of 4.
Combustion liner 530 is of conventional construction with a circumferential chamber defining a plurality of ports. The liner 530 includes a primary air combustion port 532 and a dilution air port 534 in accordance with known practices. The fuel is distributed between the combustion chamber rear wall plate 508 and the combustion liner 5.
The combustion chamber is supplied by a fuel injection nozzle 536 extending through the upstream wall 538 of 30. Igniter 54
0 is housed within the combustion liner 530 through the upstream wall 538 and the combustion chamber rear wall plate 508. Combustion liner 530 includes a port 542 that communicates with a turbine inlet area 544 formed between turbine aft wall 526 and a laterally extending portion of turbine exhaust nozzle 514 .

Turbine inlet zone control 546 is used in turbojet 500 illustrated in FIG. 24 and is the same as turbine inlet zone control 228 illustrated and described for the turbochargers of FIGS. 1-16. In particular, the turbine inlet zone controller 546
includes movable nozzle vanes 548 located circumferentially about the nozzle area and pivotable to vary the velocity and angle of the flow of combustion gases through the nozzle inlet area 544. Movable nozzle blade 548
are trunnions 55 extending from opposite sides thereof.
0,552. A trunnion 550 extends through a laterally extending portion of the turbine exhaust nozzle 514 and a trunnion 552 extends through the turbine rear wall 526 and is attached to the drive lever 554. The end of the drive lever 554 is associated with the outer race of a ball bearing assembly 556 and the inner race of the ball bearing assembly 556 is secured to the compressor rear wall 524 as described for the turbocharger 20 illustrated in FIGS. 1-16. It has been done. The action of drive lever 554 is the same as that described for turbocharger 20 of FIGS. 1-16, and controls the angular orientation of movable nozzle vane 548 as desired.

The bearing support cylindrical body 516 is connected to the support column 5 as described above.
18 concentrically within the inlet nozzle 504 . Compressor rotor 570 and turbine rotor 57
2 is a shaft 5 mounted to rotate within the bearing support cylinder 516 by two ball bearings 576 and 578.
It is supported by 74. The turbine rotor 572 is connected to the shaft 5
74 and a centrifugal compressor rotor 570 is mounted between a turbine rotor 572 and each ball bearing assembly 576,578.
The shaft 574 is connected to the turbine rear wall 526 and the compressor rear wall 5.
24. In this case, a labyrinth seal 580 formed around shaft 574 provides a seal between compressor rotor 570 and turbine rotor 572.

The turbine rotor 572 is attached to the shaft 574 by welding or the like.
It is fixed at Compressor rotor 570 is also held on shaft 574 by a retaining nut 582. Nut 582 is of the same construction as nut 180 described for the embodiment illustrated in FIGS. 1-16. Compressor rotor 570 is connected to nut 582 and shaft 57
4 and 586.

As shown in FIG. 24, the support ring 590 is mounted within the end of the cylindrical body 516 adjacent to the compressor rotor 570, and is prevented from moving within the cylindrical body 516 by a retaining ring 592 attached to the cylindrical body 516. be. Outer race 59 of ball bearing assembly 578
4 is formed within the support ring 590. inner race 5
96 is formed integrally with the shaft 574. ball 598
are received between inner race 596 and outer race 594 to form ball bearing assembly 578.

The bearing support cylinder 516 is mounted within the inlet nozzle 504 by a plurality of struts 518 extending from the inner wall 520 of the inlet nozzle 504 . This structure is similar to the structure illustrated in FIGS. 1 to 16 with respect to the bearing support cylinder 60 and its attachment within the inlet of the turbocharger 20. Cap 522 is attached to the end of bearing support cylinder 516.

As shown in FIG. 24, the compressor rear wall 524 and the turbine rear wall 526 are located within the compressor burner housing 506 and are attached to each other by threaded pieces 527.

Ball bearing assembly 576 includes an inner race 600 integrally formed with shaft 574 and an outer ring 602 slidable within cylinder 516 and having outer race 604 formed therein to receive balls 608.
Compression spring 610 includes outer ring 602 and cylindrical body 516
Fixing the position of shaft 574 by forcing outer ring 602 outwardly in conjunction with retaining ring 612 fixed therein and fixing the position of balls 606, 598 within each ball bearing assembly 576, 578, respectively. do.

The structure of the bearing for the turbojet 500 is similar to that of the turbocharger 20 illustrated in FIGS. 1 to 16.
The structure is the same as that used in . The structure and operation of each ball bearing assembly is therefore the same as described above for turbocharger 20. Further, the wick material that contacts the ramp 614 as a part for lubricating each ball bearing is the same as that illustrated for the turbocharger 20 illustrated and described in FIGS. 1 to 16.

In operation of the turbojet 500 illustrated in FIG. 24, air enters the tubular air inlet through its inlet 620 and follows the air path indicated by arrow 622. This air is transferred to the compressor rotor 570
from which it is compressed and discharged radially along the air path indicated by arrow 624. This compressed air is directed into the combustion chamber section 626 and into the combustion liner 530 through ports 532 and 534. The air mixes with fuel injected into combustion liner 530 through nozzle 536. In this way a stoichiometric mixture is obtained. This mixture is ignited within combustion liner 530 by igniter 540 . Combustion liner 53
Cooling air is introduced into the turbine through port 534 to cool the combustion exhaust gases before they enter the turbine inlet area 544. These gases are associated with turbine rotor 572 and drive rotor 572 and a compressor rotor 570 attached thereto by shaft 574.
The combustion gases are then discharged to the outside air through nozzle 514 to generate thrust by turbojet 500.

As is clear from the turbojet 500 illustrated in FIG. 24, this device is compact in size and has an extremely small number of parts. Additionally, the bearings on which the turbine and compressor rotor shafts rotate are located away from the extremely hot combustion gases that are delivered from the turbine rotor 572 and from the combustion liner 530 to the rotor 572.

Furthermore, the severe problems caused by the extremely high temperatures normally experienced by such bearing structures are usually completely eliminated by the unique bearing arrangement described above. Effective operation of the turbocharger is limited because the bearings are protected from the extremely high temperatures associated with the combustion gases used to drive the turbine and are also effectively insulated from these high temperatures by the compressor rotor and compressor rear wall and the turbine rear wall. lubrication is sufficient. This eliminates the need for a constant flow of lubrication to the bearings, as well as the many components necessary to provide such lubrication that must be driven by the rotating shaft.

FIG. 25 shows a turbo blower 650. The structure of the turbo blower 650 is the same as that of the turbojet 500 illustrated in FIG. 24, except that a bypass air flow path is provided in addition to the primary air flow path, and an axial stage compressor is provided upstream of the centrifugal compressor. is almost the same.

As shown in FIG. 25, the turbo blower 650 is
It includes an inlet nozzle 652, a main housing 654, and a combustion chamber rear wall 656. The combustion chamber rear wall 656 is attached to the main housing 654 by suitable bolts 658. A turbine exhaust nozzle 660 is attached to the combustion chamber rear wall 656 by bolts 662. A compressor back wall 664 and a turbine back wall 666 are located within the main housing 654. Combustion liner 668 is then mounted within combustion chamber 670 with a port 672 opening between turbine rear wall 666 and a laterally extending portion 674 of turbine exhaust nozzle 660 . Combustion liner 668 has a plurality of ports 676 formed therein. A fuel injection unit 678 is housed within liner 668 through combustion chamber rear wall 656 . An igniter 680 is also contained within liner 668 and ignites the fuel-air mixture during operation of turbo blower 650.

A fixed stationary vane 682 is attached between the main housing 654 and the compressor rear wall 664 by suitable threaded fittings. A bearing support cylinder 684 is mounted within the inlet nozzle 652 by a plurality of struts 686. A plurality of stationary vanes 688 are also mounted on the inlet nozzle 652 for cooperation with the axial stage compressor, as will be described in more detail below.

Shaft 690 is supported within bearing support cylinder 684 by bearing assemblies 692, 694. Each bearing assembly 692, 694 is the same as described for turbojet 500 of FIG. 24 and the turbocharger illustrated in FIGS. 1-16. An oil chamber 696 is formed between the inner wall 698 of the inflow nozzle 652 and the bearing support cylindrical body 684.
A suitable ring 700 having O-rings 702, 704 is located within the mouth of chamber 696 and is retained by ring 706. Each lamp core 708, 710 communicates through an oil chamber 696 and sends oil to ramps 712, 714, respectively, so that the bearing 692, as described for the embodiments of FIGS. 1-16 and FIG. 69
Lubricate 4.

Radial compressor rotor 716 and turbine rotor 718 are attached to shaft 690. The same turbine inlet zone controller 720 as described with respect to FIGS. 1-16 and 24 cooperates to control the velocity and angle of the combustion gases to the turbine rotor 718.

A bypass flow path 722 is formed between an inner bypass wall 724 and an outer bypass wall 726. The outer bypass wall 726 is attached to the inner bypass wall 724 by a plurality of fixed stationary vanes 728 positioned between the walls. The bypass outer wall 726 is connected to the main housing 65 by a plurality of struts 730 spaced primarily circumferentially around the main housing 654.
The position has been determined from 4.

Axial stage compressor rotor 732 is attached to shaft 690 by a suitable nut 734 that abuts rotor 732 against retaining ring 736 . The rotor 732 is
It includes a boss 738 and a plurality of blades 740 extending radially from the boss 738. A plurality of cylindrical armatures 742 are embedded within and adjacent the boss 738 of the rotor 732 and extend from the bypass wall 726 to the strut 7.
It cooperates with a fixed field winding 744 attached by 46 to produce an electric current.

The action of the turbo blower 650 illustrated in FIG. 25 is such that the air entering the inlet of the turbo blower 650 through the axial stage compressor rotor 732 is divided into a bypass flow shown by arrow 747 and a primary flow shown by arrow 748. Except for the turbojet 5 illustrated in FIG.
The effect is almost the same as that of 00. Bypass air is compressed and discharged through bypass nozzle 749 to generate thrust from turbo blower 650. The primary air is directed along a path indicated by arrow 748;
The liner 668 mixes with the fuel in the combustion liner 668.
The compressor is compressed by an axial stage compressor rotor 732 and a centrifugal compressor rotor 716 before being ignited within. The ignited gas flows past turbine inlet zone controller 720 and toward turbine rotor 718 driving turbine rotor 718 as well as axial stage compressor rotor 732 and centrifugal compressor rotor 716 . Exhaust gas from turbine rotor 718 is discharged through turbine exhaust nozzle 660 to provide additional thrust to turbo blower 650.

Turbine inlet zone control system 720 is the same as illustrated and described with respect to turbocharger 20 illustrated in FIGS. 1-16 and turbojet 500 of FIG. Controller 720 operates to control both the velocity and angle of the combustion gases that are directed from combustion liner 668 to turbine rotor 718 . Of course, the monitoring devices described with respect to FIGS. 6 and 7 with respect to turbocharger 20 apply directly to turbojet 500 and turbo blower 650 of FIGS. 24 and 25, respectively. That is, the speed of the turbine and associated compressor is monitored and the turbine nozzle inlet area is varied according to this speed or other engine parameters as desired.

While the axial stage compressor rotor 732 is rotating, the armature 7
42 rotates relative to winding 744 and generates electricity.
This electricity operates various types of engine or aircraft equipment as desired, making it a very economical and compact energy package for this equipment.

26 and 27 show a turbojet 750 as a modification of the turbojet 500 illustrated in FIG. 24. turbojet 750
is the inflow cylinder 7 joined to the main housing 754
52, and a rear plate 756 is attached to the main housing 754 with a plurality of bolts 758. A turbine exhaust nozzle 760 is attached to the end of the rear plate 756 opposite the main housing 754 by suitable bolts 762. The turbine exhaust nozzle 760 is
A laterally extending portion 766 is formed to join the ejection cylinder 764 and the main housing 754.

A bearing support cylinder 768 is supported from the inlet cylinder 752 by a plurality of fixed stationary vanes 770 .
Shaft 772 is supported within cylinder 768 by bearing assemblies 774 and 776. Each bearing assembly 774, 776 is identical in construction and operation to bearings 576, 578 illustrated for the embodiment of FIG. A turbine rotor 778 is integrally formed with shaft 772. A compressor rotor 780 is mounted on shaft 772 between turbine rotor 778 and each bearing 774,776.

A compressor back wall 782 and a turbine back wall 784 are mounted within the main housing 754 . A plurality of stationary vanes 786 are attached between main housing 754 and compressor rear wall 782 by threaded pieces 788 . A suitable igniter 789 is the turbine nozzle 760
The turbine is mounted in a chamber formed within a laterally extending section 766 between section 766 and a turbine rear wall 784 .

A turbine nozzle inlet zone control device 790 is provided that includes a movable vane 792 pivotally supported on respective trunnions 794, 796 within a laterally extending portion 766 and a turbine rear wall 784, respectively. Turbine nozzle inlet area control system 790 is the same as described and illustrated with respect to the construction of turbocharger 20 of FIGS. 1-16 and turbojet 500 of FIG. 24. Turbine rear wall 78
A labyrinth seal 800 is formed within 4 to effectively seal between the turbine section and the compressor section. Compressor rotor 780 is attached to shaft 772 by nut 802 which abuts rotor 780 against stage 804 of shaft 772.
It is attached to 72.

Axial compressor 806 rotates with shaft 772 and cooperates with stationary vanes 770 to compress air directed into compressor inlet 808 formed by inlet cylinder 752 . In the embodiment illustrated in FIGS. 26 and 27, an axial hole 812 is formed in the shaft 772. A fuel supply pipe 814 is connected to the hole 812 and supplies fuel into the axial hole 812 . A suitable seal 816 is positioned between the fuel supply tube 814 and the surface of the bore 812 to prevent loss of fuel at the connection between the supply tube 814 and the shaft 772.

A plurality of radial ports 818 extend through shaft 772.
is formed to provide communication between the hole 812 and an annular groove 820 formed in the compressor rotor 780. A plurality of radial ports 822 communicate from groove 820 through compressor rotor 780. O-shaped ring 826, 828
A suitable seal, such as , is mounted within an annular groove in rotor 780 to prevent fuel leakage at the connection between rotor 780 and shaft 772.

During operation of the turbojet 750 illustrated in FIGS. 26 and 27, fuel is pumped through the hole 812 of the shaft
is supplied through a fuel supply pipe 814. A fuel supply path is formed between hole 812 and the air flow path indicated by arrow 810 (FIG. 26) via port 818, groove 820, and port 822. This fuel flow path is indicated by arrow 838. When the radial compressor rotor 780 is rotated, the fuel flows through the holes 81 due to centrifugal force.
2 through this fuel flow path into the air path indicated by arrow 810. At the same time, the air
The air is sucked into an inlet 808 by an axial stage compressor 806 and delivered in a compressed state along an air path. This air is further compressed by radial compressor rotor 780 and mixed with fuel downstream of rotor 780. Fuel is transferred from port 822 to radial compressor rotor 7
The action of the compressor rotor blades on this fuel causes it to be highly atomized as it is discharged through the air pump 80. This atomized fuel, when mixed with compressed air fed into turbojet 750, is directed along the air path and ignited by igniter 836 upstream of turbine rotor 778. The exhaust resulting from combustion passes the blades of turbine rotor 778 and drives the rotor in a conventional manner. These exhausts are then discharged from nozzle 7.
60 and generates thrust from the turbojet 750. Of course, combustion temperatures are significantly higher with this combustion method and materials that can withstand these higher temperatures are required.

The apparatus of FIGS. 25 and 26 eliminates many of the parts previously required in a conventional turbojet. The unique combustion delivery system illustrated in FIGS. 25 and 26 eliminates the need for a fuel pump and all associated hardware. Instead of an ordinary fuel pump, it automatically releases fuel in atomized form by injecting it into the air stream compressed by this rotor through a hole in the compressor rotor shaft and through ports on the compressor rotor itself. It uses a fuel system that 25th
The apparatus shown in the figures and FIG. 26 has a two-stage compression structure in which an axial-flow stage compressor is added upstream of a radial-flow compressor, that is, a centrifugal compressor. This construction provides additional balance to the system by providing the axial stage compressor on the side of the bearing opposite to the centrifugal compressor side.

Additionally, this design positions the bearing immediately upstream of the turbine rotor and away from the turbine rotor and combustion chamber. This construction insulates the bearing from the extremely high temperatures experienced by this area of the device. Therefore, there is no need for an oil device. Furthermore, the device is suitable for very small turbojet or turbo blowers.

FIG. 28 shows a turbo blower 900 that is a modification of the turbojet 750 illustrated in FIG.
The structure of the turbo blower 900 is similar to that of the second air blower except that a bypass air flow path is provided in addition to the primary air flow path.
This is almost the same as the turbojet 750 illustrated in FIG. Turbo blower 900 includes an inlet cylinder 902 joined to a body 904 and a rear wall 906 attached to main housing 904 by bolts 908 . Turbine discharge nozzle 910 is bolted to 91
A laterally extending portion 913 of the nozzle 910 supported from the rear wall 906 by the main housing 904
is related to. A bearing support cylinder 914 is supported from the inlet cylinder 902 by a plurality of struts 916 . A compressor back wall 918 and a turbine back wall 920 are supported within the main housing 904 . Turbine inlet zone controller 921 same as described for controller 790 of turbojet 750 of FIG.
is provided. The shaft 922 is connected to each bearing assembly 92.
3,924 and is rotatably supported within the bearing support cylinder 914. Each bearing assembly 92
3,924 is the same as the bearing described for the turbojet 750 illustrated and described in FIG. A turbine rotor 926 is attached to one end of shaft 922. A compressor rotor 928 is also mounted between the turbine rotor 926 and each bearing assembly 923,924.

In FIG. 28, an axial compressor rotor 930
is attached to the end of the shaft 922 and the inflow cylinder 9
02 and a secondary flow path formed between the secondary flow outer wall 932 and the secondary flow inner wall 934. Secondary flow outer wall 932 is inner wall 93
4 to a plurality of stationary vanes 936.
Air directed through the secondary flow path as shown by arrow 938 is directed to compressor rotor 928 and vane 9.
36 and is discharged through the discharge nozzle 940 to generate thrust from the turbojet.
Air directed along the primary flow path as indicated by arrow 942 is compressed by compressor rotor 928 and mixed with the fuel provided in the air stream in the same manner as described for the embodiment of FIG. igniter 946
ignites. The combustion gases are directed to the turbine rotor 9 via a turbine inlet zone controller 921.
26 and compressor rotors 930, 928. Air exiting past turbine rotor 926 and through turbine exhaust nozzle 910 produces additional thrust from this turbofan.

In FIG. 29, a generator 970 is mounted within the compressor inlet nozzle. As shown in FIG. 29, a generator 970 includes an armature 972 attached to the end of a compressor and turbine 974. As shown in FIG. 29, shaft 974 is fitted with bearings 976 and a second bearing (not shown) which are the same as those shown for the turbocharger of FIGS. 1-16 and the turbojet and turbo blower of FIGS. 24-28. It is supported by Field winding 978 is mounted within bearing support housing 980. Housing 980 is mounted concentrically within compressor inlet nozzle 982 by vanes 984 in the same manner as described above for the turbocharger, turbojet, and turboblower embodiments. A cap 986 is attached to the end of the bearing support housing 980.

During operation of generator 970, armature 972 rotates with shaft 974 within field winding 978 to produce electrical current in well-known manner. The resulting current is directed from field winding 978 by conductors (not shown) to any location within the device requiring power. That is, when the device shown in FIG. 29 is used in a turbocharger, the generator is used to supply current to any part of the vehicle or other device requiring electrical power. When the structure described in FIG. 29 is applied to the turbojet of FIGS. 24-28, electrical energy can be used to power the induction system or other components that require electrical power.

That is, as shown in FIG. 29, an extremely simple generator can be obtained that is directly driven by the rotation of the shafts of the compressor and turbine. Furthermore, this location of the generator results in a very simple device, which is light in weight and does not interfere with the operation of the power-inducing device.

This results in a turbojet and turboblower structure that separates the combustion chamber from the bearings used to support the compressor and turbine shafts. An axial compressor is also provided on the opposite side of the bearing from the centrifugal compressor. Furthermore, fuel can be introduced into the turbojet system by supplying fuel directly from the compressor rotor shaft through the compressor rotor. In this case, fuel is released by centrifugal force and there is no fuel pump or associated hardware.

A bypass flow and a primary flow are created in the turbocharger forming the turbo blower.
In this case, an axial stage compressor is provided upstream of each bearing on the opposite side of the centrifugal compressor.

A generator is also provided with an armature mounted at the end of the compressor and turbine shafts on opposite sides of the compressor and turbine in each bearing, and with a field winding mounted within the bearing support. This construction provides a very simple generator to power the entire device.

Although preferred embodiments of the invention have been described in the foregoing detailed description and illustrated in the accompanying drawings, the invention is not limited to the embodiments described above and may be modified in many ways without departing from the spirit of the invention. Of course, each part and each element can be changed. Accordingly, the invention is susceptible to such changes and modifications and replacement of parts and elements within the scope of the appended claims.

[Brief explanation of drawings]

Fig. 1 is a perspective view of a turbocharger equipped with a bearing support structure of the present invention, Fig. 2 is a vertical sectional view taken along line 2-2 in Fig. 1, and Fig. 3 is an embodiment of the bearing support structure of the present invention. Fig. 4 is an enlarged vertical sectional view of Fig. 4-4 of Fig. 2 with the rear wall of the compressor removed for simplicity.
FIG. 5 is a partially cutaway end view taken along line 5-5 when viewed in the direction of the arrow in FIG. 2, and FIG. FIG. 6 is a partially cutaway end view of the turbocharger shown in FIG. 1 employing another control method for the turbine inlet vanes, and FIG. 6A is a block configuration diagram for explaining the control device of FIG. 6. , Figure 7 is 7- in Figure 6.
Vertical sectional view along line 7, Figure 8 shows the turbine rotor.
Compressor - An exploded perspective view showing the turbine shaft, compressor rear wall, turbine rear wall, compressor bushing, compressor rotor, and retaining sleeve.
10 shows the use of the spring between the retaining sleeve and the compressor rotor, and FIG. 11 shows an exploded perspective view of the two components of the compressor housing making up this component. Figure 12 shows the housing separated from the mold used for
13 is a sectional view taken along line 13-13 in FIG. 12, FIG. 14 is a sectional view taken along line 14-14 in FIG. 12, and FIG. is a sectional view taken along the line 15-15 in FIG. 12, FIG. 16 is a diagram showing the front diffuser wall modified to change the characteristics of the turbocharger, and FIG. 17 is a diagram illustrating the performance of a compressor for a conventional turbocharger. Figure 18 is from Figure 1 or 1.
Fig. 6 is an explanatory diagram of compressor performance for the turbocharger shown in Fig. 6, Fig. 19 is a vertical cross-sectional view of a modified version of the turbocharger shown in Figs.
21 is a top plan view of FIG. 20, and FIG. 22 is a vertical sectional view showing the turbocharger shown in FIGS. 1 to 16 installed on an in-line engine. 23 is a top plan view of FIG. 22, FIG. 24 is a vertical sectional view of the upper half of the turbojet equipped with the bearing support structure of the present invention, and FIG. 25 is a vertical sectional view of the upper half of the turbojet equipped with the bearing support structure of the present invention. 26 is a vertical sectional view of the upper half of the modified embodiment of the turbojet shown in FIG. 24, FIG. 27 is a sectional view taken along line 27-27 in FIG. 26, and FIG. The figure is a vertical sectional view showing the upper half of a turbofan equipped with the bearing support structure of the present invention, and FIG.
FIG. 6 is a vertical sectional view of a generator that is applied to the turbocharger shown in FIG. 6 and can also be applied to the turbojet and turbofan shown in FIGS. 24 to 29. 60...Outer housing (bearing support tube), 16
0...Axis, 200, 204...Fixed race, 20
6...First inner lace, 208,216...Ball,
210...Second inner race, 212,214...
Outer race ring, 222, 224...wick, 2
26... Peripheral slope part (slinger).

Claims (1)

[Scope of Claims] 1. A bearing support structure for supporting rotation of a compressor and a turbine of a turbocharger, comprising: (a) a first bearing support structure extending from the compressor and the turbine;
and (b) an outer housing having a fixed race corresponding to the inner race so as to receive a plurality of balls between the inner race and the first inner race. (c) a second outer race ring slidable relative to the fixed race of the outer housing and disposed adjacent to the second inner race to receive a plurality of balls therebetween; and (d) by arranging the ball between the fixed race of the outer housing, the first inner race, the second outer race ring, and the second inner race, (e) spring means formed on the shaft and biasing the second outer race ring away from the fixed race of the outer housing to position the shaft relative to the outer housing; (f) a peripheral ramp angled toward said ball disposed between a fixed race of said outer housing and said first inner race; The lubricant is applied to the peripheral slope and carried to the ball by centrifugal force through a wick having one end in contact with the lubricant and an opposite end in contact with the peripheral slope. lubricant means for supplying a lubricant to the peripheral slope. 2. Claims in which the lubricant means lubricates the balls between the fixed race of the outer housing and the first inner race by spraying a lubricant thereon. The bearing support structure according to item 1. 3 a second peripheral ramp formed on the shaft and positioned between the second outer race ring and the second inner race and angled toward the ball; While the shaft is rotating, the lubricant is
one end in contact with the lubricant so that it is carried to the ball by centrifugal force applied to the peripheral slope of the lubricant;
1. Lubricant means for supplying lubricant to the second peripheral ramp via a wick having an opposite end in contact with the second peripheral ramp. Bearing support structure as described.