JPH02218816A - Bearing structure to support compressor of turbo charger and revolution of turbine - Google Patents

Abstract

PURPOSE: To improve lubricating performance and improve efficiency of a compressor by providing a common shaft of a turbine and a compressor with two inner races, and disposing an outer race corresponding to the inner race in an outer ring which is slidably mounted in an outer housing. CONSTITUTION: In a structure where a shaft 160, which directly couples a turbine rotor 170 and a compressor rotor 172 of a turbocharger, is supported by two ball bearings 162, 164 in a bearing support tube 60. Each of the ball bearings 162, 164 comprises inner races 210, 206 formed integrally with the shaft 160; and outer races 214, 204 engaged with the inner races 210, 206 via a plurality of balls 216, 208 in such a manner that they can roll freely. The outer race 214 is formed in an outer ring 212 which can freely slide in the bearing support tube 60, and the outer ring 212 is biased outward by a spring 218 to leave far away from the outer race 204. Further, due to a centrifugal force added to a peripheral slant portion provided to the shaft 160, a lubricating agent in wicks 222, 224 is absorbed to be supplied to the peripheral slant portion.

Description

DETAILED DESCRIPTION OF THE INVENTION Technical Field The present invention relates to a bearing support structure that supports the entire rotation of a compressor and a turbine of a turbocharger.

REFERENCES TO RELATED APPLICATIONS This application is filed in US Pat.

BACKGROUND OF THE INVENTION Turbocharging is a means of greatly expanding the power range and susceptibility of internal combustion engines, and has become an accepted technology in practice, often in applications with engine speeds of 200 horsepower or more. This is necessary for powerful diesel engines.

Turboteering is also used, for example, in aircraft engines to maintain power when eaves are increased.

Effective application of turbochargers to internal combustion engines typically increases power output e by 50-100 percent and reduces full load specific fuel consumption (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 optically high, then the positive pumping loop will add to the net cycle output.
1oop ) will occur.

Turbocharging typically reduces full-load specific fuel economy;
It is considered a means of increasing horsepower.

With the development of engines that currently utilize turbochargers, such as large powerful diesel engines, aircraft engines, competitive motor vehicles, and similar engines, the foregoing understanding is valid and has limited application. In this example, the engine is operated near opening load or full load for most of the work cycle. However, in many applications, it is not necessary for these engines to operate at or near full load for extended periods of time. In many practical applications, such engines are typically operated at less than 50% power;
It operates with the following power. Application examples include automobiles,
Light or medium trucks, generator tops, compressors, tractors,
This is an engine used for building equipment and equipment similar to strainers.

Engines that operate in low power environments, such as those found in these engines, 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 a pump engine, this inefficiency results from increased pumping loop losses as the load decreases. This is because friction in an internal combustion engine remains relatively constant regardless of load.

Therefore, - by using a smaller and more efficient engine size, i.e. by using an engine with a smaller displacement ratio (1/2 displacement times the engine speed for a four-stroke engine), 71. Reduction in displacement, reduction in operating speed, or By the combination of two swords such as υ, part load fuel consumption can be totally improved. In gasoline engines, this improvement occurs as a result of reduced engine friction and reduced pumping loop losses. In diesel engines, this improvement is the result of - higher thermal efficiency due to higher combustion temperatures and reduced engine friction. In many applications, turbocharging is −
layers may be used to allow the use of engines with smaller effective engine sizes.

By performing a turbocharging song, it is easy to obtain naturally aspirated power that is twice the displacement per cubic inch, or in some cases six times as much. However, attempts to fully turbocharge smaller engines have generally been unsuccessful. This failure is due to current designs that allow turbochargers to be assembled around journal bearings and flat disk (7pf3) thrust bearings. This type of bearing system requires a total of 1 to 6 horsepower (depending on the specific turbocharger and the speed required in the application) just to overcome friction. This loss means that the turbo/f! puts out over 60 horsepower during the compression process! : Not important in the required application (typically a 200 hp or so engine), but very important when fully turbocharging an engine of 100 hp or less. For example, turbocharger turbines. For power outputs of 60-80 hp, bearing friction losses of 2-3 hp are not significant.However, in smaller turbochargers where the turbine power is only 15 hp - a smaller turbocharger is operated. For higher rpm, bearing friction losses of 2-3 hp or even 4-5 hp may be
This corresponds to almost 1/3 of the total turbine horsepower generated, a totally unacceptable loss.

Also, the bearing systems currently in use require significant radial and axial clearance to accommodate oil flow and rotor stability. Such clearances can affect the efficiency of both the compressor and turbine by extending the relatively large clearances associated with the plating of the compressor and turbine rotors. For example, Schnall 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 surrounding structural members. If the blade height is 1 inch, the ratio of clearance to blade height is only 11/2 percent. However, if a smaller turbocharger with a blade height of, say, 0.2 inches is desired, then
A 0.015 inch clearance between the vane and its surrounding structural members amounts to 7-172 percent of the vane height.

Therefore, while a 0.015 inch clearance may be acceptable in larger turbocharger applications, it becomes completely unacceptable when a smaller turbocharger is 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 during startup or contamination of the engine oil. When high-speed Schernal bearings are used in conventional turbochargers, continuous oil flow is inherently required to stabilize the shaft and to carry away the heat generated by viscous friction. nru. The oil flow is also used to carry away heat transferred to the bearing system from the adjacent turbine (which operates at temperatures as high as 1600°F').
is required. Even if anti-friction ball bearings are used instead of journal bearings in conventional turbochargers,
Continuous oil flow is required to carry away all the heat transferred from the turbine. Therefore, although modern turbochargers require lubrication to operate properly, lubrication is also the cause of many failures. Additionally, 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 i''LL passing through the turbine.

There are currently basically two methods used to control the power output of a turbine. The first method is to carefully size the turbine and turbine nozzle so that at maximum engine operating speed and load, the desired boost pressure is below a predetermined limit at full speed. The disadvantage of this method is that the available boost pressure is limited at low engine speeds and the response to demand is slow. The second method used to control boost pressure through the turbine is to create excess turbine power at maximum engine speed and load - dimensioning the bin nozzle and exhaust relief valves. (wastegat
e) f! : To use. In this method, when a predetermined boost pressure is reached, an exhaust relief valve opens to partially and completely bypass the exhaust gas. This method is -
Although it increases the available boost pressure at lower engine speeds and improves demand response, it is completely inefficient in that bypassed high pressure exhaust is simply wasted at the expense of increased engine back pressure. Furthermore, at part load, when the turbocharger is essentially not operating, the small nozzle area acts as a restraint on exhaust, causing increased pumping loop losses.

Therefore, a need has arisen for a turbocharger that can be effectively operated to fully turbocharge both large and small internal combustion engines. Continuous oil flow eliminates the problems previously experienced with lubricated bearings in turbocharger turbines. A turbocharger with a bearing system that effectively uses the entire flow of propellant gas is required.
I started to do that. In addition, the bearing assemblies for the compressor and turbine rotors are fully integrated into the desired compressor and turbine rotors.
It must be possible to easily reduce the rotor clearance.

The construction and design of conventional compressor fittings for turbochargers also presents a considerable 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 and flows toward the compressor rotor.
It is accelerated to near the speed of sound by using a metal plate at a right angle to the inlet channel. This is accomplished by increasing the air pressure and reducing the total velocity of the accelerated gases released from the tips of the compression rotor blades. This process, known as diffusion, is accomplished more effectively by reducing the velocity of the gas without creating turbulence, whereby a large percentage of the velocity energy is converted to pressure energy, reducing the static pressure Raise it completely.

To facilitate this diffusion process, turbochargers using centrifugal compressors use a compressor rotor blade that follows the profile of the rotor blades from the leading edge to the outer tip. Usually has walls. The compressor rotor wall then extends past the outer tips of the rotor blades;
Form one layer of the two walls of the diffuser and then end to form a peripheral gap. The compressed gas f' is
This wall, which flows through this peripheral gap into the peripheral chamber leading to the engine suction manifold V, faces the compressor rotor blades, contours closely to the compressor rotor blades, and then extends outwardly allows the compressed gas to flow into the peripheral chamber. The velocity of the compressed gas is uniformly reduced after it leaves the compressor rotor blades and before it enters the peripheral chamber leading to the engine. This wall structure therefore greatly increases the static pressure generated by the compressor.

To create this structure, many turbocharger compressor housings are sand cast (θand-cast) so that the outer circumference of the compressor--the outer periphery of the compressor--and the compressor wall are one cast member. . Usually this was accomplished by using a 5 and core to form the entire peripheral chamber leading to the engine's suction manifold. After casting, this sand core is removed to form a peripheral chamber on the wall opposite the compressor, into which gas flows from the tip of the compressor rotor.
leave.

Although die-casting of the compressor housing is substantially less expensive and more precise than sand-casting, it is possible to form a peripheral chamber through which the compressed gas flows into the engine's suction manifold, while at the same time forming the entire diffuser wall. It has not been possible to use die castings of optimal design because the entire die casting mold cannot be used for this purpose.The variable area chamber is necessarily larger than the inlet gap, through which the gas Injected from compressor rotor blades
Because of this, it was not possible to die-cast an optimally designed compressor housing. This is because it is not possible to design a complete mold that can form such a passageway behind the wall facing the compressor rotor blades.

If a die-cast compressor housing is used, the molds must be combined or split to make the casting, since the walls normally formed in the sand-cast compressor housing are simply enlarged or removed. Can be done. However, without this wall, the gases accelerated by the compressor rotor would prematurely pass from the diffuser into the surrounding chamber leading to the engine intake manifold.

As a result, compressor housings of such construction substantially reduce compressor efficiency and therefore compressor performance.

DISCLOSURE OF THE INVENTION According to the present invention, an anti-friction ball bearing assembly is used for full shaft support of a compressor turbine.

The use of ball bearings is particularly important to provide better control over the radial and axial movement of the turbine and compressor. This additional control over the movement of the rotating bearing assembly improves overall efficiency of the compressor and turbine by allowing for a reduction in 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 mounted to the turbocharger housing, and a plurality of balls are receptacles between the fixed outer race and the first inner race. 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 nosing and the second outer race ring to move the outer race ring into the second outer race ring. Shift this second outer race away from the first outer race.
It is associated with a ball located between the race ring and the second inner race. At the same time, a first outer race is added to the turbocharger housing and is associated with a ball between this and a first inner race on the shaft of the turbine and compressor. nru.

DESCRIPTION OF THE PREFERRED EMBODIMENTS FIG. 1 is a perspective view of a turbocharger 20 having a bearing support structure according to 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 la/zo band 28.

As shown in FIGS. 1 and 2, the compressor housing 24 includes a cylindrical inlet hole 40 having a lateral wall 42 added to one end thereof and extending outwardly therefrom. A circumferential chamber 44 is added to the wall 42. The inlet hole 40 defines the entire compressor air inlet 50, and the chamber 44 defines the compressor outlet 52. Turbine housing 26 defines a turbine air booster 54 and a turbine exhaust 56 .

In use, the turbocharger draws air into the inlet 50 and exhausts the compressed air through the exhaust port 52 to the internal combustion engine to which the turbocharger is attached. Exhaust air coming from the engine is introduced into a turbine air inlet 54 to fully drive the turbine for the turbo engine, and is discharged through a turbine exhaust port 56.

A bearing support cylinder 60 is augmented 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 added on the end of this bearing support cylinder 60. A piston type actuator or actuator 80 is added to the turbine housing 26 by means of a bracket 82. The actuator 80 includes a controller 84 that is operated to extend and retract a control rod 86 as described below. To operate the control rods 86, air lines 88, 89 are provided to feed the air controller 84 as required. Compressor - Add an oil sump cover plate 9 (1) to the Usuong 24 with a plurality of screws 92.

As shown in FIG. 2, the compressor rear wall 100 and the turbine rear wall 102 are positioned intermediate the compressor housing 24 and the turbine housing 26 when both housings are assembled. . These four components are assembled by a V-shaped Lanzo band 28 with mutual guidance. As can be seen from a closer look at FIG. 2, only a single V-shaped lamp is required to hold the entire assembly together. As previously mentioned, the compressor housing 24 is provided with a cylindrical inlet hole 40' having a lateral wall 42 extending outwardly therefrom at one end thereof. 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 and is directed toward discharge from the compressor outlet 52. increase

The inlet hole 40 has a first interior wall 110 with a diameter that tapers toward the wall 42 and a second interior wall 112 connected to the first interior wall 110 by a step 114. have. This second wall portion 112 has a diameter that widens towards the end of the 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.

Front compressor wall insert) 126fl, cylindrical throat 12
8 and a circular disk 130 mounted laterally from one end of the throat 128.

The throat 128 has 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 . The widening diameter retaining surface 132 corresponds to the widening surface of the inner wall 112 of the entrance hole 40 so that the throat 128 mates with the throat inserted into the entrance hole 40. Further, the wall surface 131Jd of the throat portion 128 having a smaller diameter corresponds to an extension of the smaller diameter of the first inner wall portion 110 of the inlet hole 40. When the insert 126 mates with the inlet hole 40, the inlet hole 400A has a diameter that decreases from the mouth into the turbocharger.

A plurality of rivet-like protrusions 140 extend from disk 130 and correspond to holes 116 in wall 42. An insert 126 is associated with the compressor housing 24 and also connects the throat 12.
If the end of 8 is associated with the step 114 of the inlet hole 40, the protrusion 140 is associated with a disc 130 which is drilled into the hole 116 and abuts a corresponding surface of the wall 42. As shown in FIG. 2, protrusion 140 is inserted into hole 116 and its head is modified to attach insert 126 to housing 24. As shown in FIG. The disk 130 extends beyond the wall 42 to form the chamber 4.
4 so as to partially cover the hole 118. disc 1
A circumferential gap 146 between the outer edge of 30 and the wall of chamber 44
, and a seven-day zone 148 is formed between the compressor rear wall 100 and the disk 130 between the centrifugal flow compressor rotor 112 and the gap 146 leading to the chamber 44 .

As shown in FIG. 2, the bearing support tube 60 is supported concentrically within the inlet hole 40 by a plurality of struts 62 extending inwardly from the wall 64 of the inlet hole 40. As shown in FIG.

The turbocharger 20 has two ball bearings 162° and 164°.
The shaft 1 is supported to rotate within the bearing support cylinder 60 by
I have 60. Radial turbine rotor 170 has shaft 1
60 and a centrifugal compressor rotor 172 is mounted intermediate the turbine rotor 170 and bearings 162,164. The shaft 160 extends through a hole 176 in the compressor rear wall 100 and a labyrinth seal 17E1 in the turbine rear wall 102.

The radial turbine rotor 170 is attached to the shaft 16 by welding or the like.
0 and the centrifugal flow compressor rotor 172 is attached to the retaining nut 1.
80 and held in position on axis 160. Centrifugal flow compressor rotor 172 is bored to receive shaft 160 and has hole 18
Counterbore to form 2. The hole 182 has a diameter larger than the outer diameter of the retaining nut 180 and
80 is pressed so as to be associated with the bottom wall 184 of the hole 182 and is held at a position above the entire compressor rotor shaft 160. A compressor rotor shim 186 is placed between the centrifugal compressor rotor 172 and the stage 188 of the shaft 160 to ensure precise axial positioning of the centrifugal compressor rotor 172'ffr:.

In the bearing support structure of the present invention, as is clear from FIGS. 2 and 6, a 200 ftn ring is installed within the end of the bearing support cylinder 60 adjacent to the centrifugal compressor rotor 172! 2 and a retaining ring 202 attached to this bearing support cylinder 60.
This prevents it from moving into the bearing support cylinder 60. Ring 2001': Outer race 2 of bearing R164
04 is formed, and the same race 206 is formed integrally with the shaft 160. Between the inner and outer races are balls 208, which form the bearing assembly 164.

The bearing assembly f162 includes an inner race 210 formed integrally with the shaft 160, and an outer ring 2 formed to receive the balls 216 and slidable within the bearing support cylinder 60 together with an outer race 214.
Equipped with 12t. Ring 212 and bearing support tube 60
A compression spring 218 is provided between the retaining rings 220 fixed therein, biasing the rings 212 outwardly to fix the position of the balls 216, 208 in the bearing arrangement 162, 164, respectively. In this manner, the position of the shaft 160 is fixed.

As shown in FIGS. 2 and 6, the outer race 204 is connected to the ring 2.
It is formed so that the radius of the ball is only on the 10,000 side. Therefore, assembly of the bearing assembly 164 is accomplished by placing the entire ball 208 within the race 206 and linking the ring 200t to the shaft. Similarly, outer race 214 is formed within ring 212 with a ball radius only on the 10,000-degree side. Bearing device 1
The 62 balls 216 are assembled by moving the outer ring 212 to compress the spring 218t and inserting the entire ball 216 into the race 214 of the shaft 160. ring 2
By releasing 12 , spring 218 automatically engages the ring with ball 216 to form bearing arrangement 162 and simultaneously associates ring 200 with ball 208 of bearing arrangement 164 . An outer race 214 formed on the ring 212 and the shaft 16
The gap between the ball 2 and the inner race 210 formed at
The diameter of ball 216 is made smaller by the amount of light, and ball 216 becomes
This prevents the cap 66 from popping out from the position shown in FIGS. 2 and 3.

Alternatively, by using a suitable retainer, the ball 20
8.216 bearing device 162.
164 can be used. Depending on the application, one may use an oil-impregnated cage or a sacrificial cage that replenishes the balls with a self-lubricating coating. Mounting of the shaft 160 within the support tube 60 is completed by linking a cap 66 to the end of the bearing support tube 60 that closes a hole in the support tube 60 remote from the compressor rotor 172.

One-piece inner race full ball: The use of a lubrication device 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 total axial or radial movements. As a result, such a bearing system greatly reduces the clearance required between the housing surrounding the compressor and turbine head, and concentricity problems are minimized. become.

In a preferred embodiment of the invention (162°16
4 is in a so-called oil-starved state. The only lubrication provided to this bearing arrangement is by wicks 222, 224, which are moved from the oil sump (R) to the ramp or slinger 226 by oil-containing capillary action. Slinger 22
Due to centrifugal force, the oil supplied to the shaft 160 is thrown onto the bearing arrangement 162, 164 during rotation of the shaft 160.

In this way, the use of machine oil as a lubricant for the turbocharger bearings and the associated pipe work and covert work are eliminated. Moreover, accidents resulting from the use of dirty machine oil as a lubricant or from the lack of machine oil during start-up are avoided. Furthermore, there is no need for oil seals, and bearing accidents caused by seal failure forces are also eliminated.

Alternatively, the bearing arrangement may be permanently lubricated by dense oil or grease packed into the race and around the balls of the bearing arrangement. Yet another method is to use an oil-impregnated phenol retainer to provide a full supply of lubricant to the ball over a fairly long period of time. In all such cases, the life of the bearing will be shortened, but the need for the wicks 222, 224 and the slinger 226 will be eliminated.

Conventional turbochargers that use journal bearings or disk-type thrust bearings require continuous lubrication of the bearings. Additionally, if the bearing is adjacent to the turbocharger, the turbine
Since the bearings are exposed to temperatures up to

Even ball bearings require a continuous flow of oil for cooling! ! - Require. The ability of the present bearing system to operate well without the need for conventional large amounts of lubrication for bearing lubrication and cooling is due to the special bearing arrangement and compressor and turbine is the result of the relative position of the bearing to .

As will be explained later, since the bearing device does not require a continuous flow of oil f'L for lubricating and cooling the bearing, the bearing support structure and the turbochanger can be added in any direction. On the other hand, conventional turbochargers have extremely strict restrictions on the orientation in which they can be used.

In the turbocharger shown in FIG. 2, the Schernall bearing and disk bearing of a conventional turbocharger are replaced by a precision pressure-type bearing arrangement, eliminating the need for a continuous flow of lubrication for the bearing. ing. Moreover, the second
The compressor and turbine in the turbocharger shown are both overflowing on one side of the bearing arrangement,
The turbine is maximally separated from the bearing arrangement by placing the compressor between the bearing arrangement and the turbine. This arrangement provides good thermal insulation between the turbine and the bearing system so that the bearing does not heat up as much as conventional lubrication would require.

The arrangement of the compressor and turbine aft together in Figure 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 bushings and provides a more compact package than current straddle-style mounted rotors with offset support bearings intermediate the compressor and turbine.

Additionally, the use of anti-friction ball bearing devices 162,164
The radial and axial movement of the compressor and turbine can be well controlled and the next end gap can also be reduced.

This significantly improves the efficiency of the compressor and turbine. Ball bearing device 16 with reduced friction compared to Schernall bearings
2.164 reduces the turbine work required to drive the lubricant bearings. This also reduces engine backpressure, resulting in improved 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. 1st, 2nd
As can be seen in the figure, the exhaust gas from the internal combustion engine to which the turbocharger is attached is transferred to the air inlet 54' of the turbine.
T is injected into the turbocharger through a nozzle area 23 formed by the turbine rear wall 102, a rock 232 parallel to this, and a nozzle vane 234 (one through which leads to the blades of the turbine rotor 170). This nozzle area is located circumferentially around the turbine rotor 1.
A structure 228 that includes a plurality of movable nozzle vanes 234 that are rotatable to vary the exhaust flow rate and angle up to 70.
control (wealth) by

21st. As shown in FIG. 4, vane 234 has a trunnion 236,238'r extending from its opposite side.

A trunnion 236 passes through the turbine rear wall, 102 and attaches to the operating lever 240. Trunnion 238 is connected to wall 23
It is postponed to within 2 days.

A nipple 242 is formed at one end of the operating lever 240 . These nipples extend into radial slots 244 formed in control ring 246. control ring 2
46 and the operating lever 240fl, located in the air gap gap 247FF3 intermediate the pressure 1 m rotor 172 and the turbine rotor 170.

The control ring 246 is located concentrically about the @ line of the ° axis 160 and extends from the cylindrical bulge 24 from the compressor rear wall 100.
Receive it on 8.

In FIG. 2, the control ring 246 is formed with an inner race and an outer race, and includes an inner ring 250 and an outer ring 252 that receive a plurality of balls 254 between them.
It is equipped with The ring 250 is fixedly attached to the cylindrical expansion surface 248 extending from the rear wall 100 of the compressor, and the outer ring 250
2, it rotates angularly with respect to the inner ring. As shown in FIG. 4, by rotating the outer ring 252, each operating lever 2
40 pivots about the axes of trunnions 236, 238, resulting in each nozzle vane 234 pivoting simultaneously. As shown in FIGS. 4 and 5, one of the operating levers 240
The lever has an extension 262 in its entirety. Control rod 86 includes a threaded eyepole 86ak attached to extension 262 by an axle pin. The opposite end of the control rod 86a is threaded within the control rod 86 and is adjustable therein to permit readjustment of the nozzle vanes by turning its axis of rotation. The movement of the control rod 86 causes the operating lever 240 to fully pivot and the outer ring 25 of the control ring 246 to rotate.
2, rotate all the way, and use this to rotate each of the other operating levers 240 and the attached nozzle blade 234.

FIG. 5 shows an actuator 80 for use in an automatic cycle engine. In this system, a control rod 86 is controlled by a piston-type actuator 80. Alternatively, these pistons may be replaced by diaphragms. Actuator 80 is controlled by compressor discharge pressure, which is sent to controller 84 through pipes 88A, 88Bk. The increased pressure entering actuator 80 causes an extension of control 886 and a corresponding opening of the compressor nozzle area.

As shown in FIG. 5a, control n86 is controlled by controller 84 to change the position of nozzle vane 234. As shown in FIG. Controller 84 includes a housing 264 forming a large cylinder 265 and a small cylinder 266. Front 268a
A large piston 268 having a rear surface 268bl and a rear surface 268bl is inserted into the cylinder 265. Front surface 270a and rear surface 2701:
A small piston 210t with +' enters the cylinder 266.

Attach the front plate 272 to the housing 264 with appropriate bolts, and provide a hole 273 for receiving the entire control rod 86 inside the housing. The piston 270il-r is fixedly attached to one end of the control rod 86, and the piston 268 is the control rod 86? #4186
The front surface 2 of the piston 268
The only stop is an additional ring 274 fixed to the control rod 86 adjacent to 68a.

A spring 2γ6 is provided in the rotation of the control rod 86, and the piston 26
8 and the front plate 272. In this case, a large spring 278 surrounds 86 control rods and is inserted between piston 270 and piston 268. 5 nitrogen trachea pipes 88B and 88A are connected to the cylinder 266 leading to the rear surface 270b of the piston 270,
The cylinder 26 also reaches the front surface 268a of the piston 268.
5, respectively. One end of the air pipes 88A, 88B is attached to the intake manifold of the engine, and the other end is attached to the controller 84, which controls the manifold pressure.
This leads to the front surface 268a and rear surface 270b of the pistons 268, 270. The air pipe 86 is exposed to the piston 2.
Full atmospheric pressure is supplied to the rear surface 268b and front surface 270a of 68,270.

An annular group 280 and a hole 282 through a guide cylinder 284 projecting from the rear wall 268b of the piston 268 facilitate the transmission of atmospheric pressure along the entire rear surface of the piston 268. Similarly, the radial grouping 286 on the rear surface of the piston 268 facilitates the transmission of pressure to the entire rear surface of the piston.

The controller 84 opens the nozzle vanes 234 to the turbine rotor 110 to both low and high manifold pressures.
c When the 6H+J control rod 86 is fully extended, the nozzle blade at intermediate manifold pressure? The plan is to retract the control rod 86 to close it. In further detail, while operating in the closed throttle position when the suction manifold is at a vacuum, such as 5 to 3 pei absolute pressure, the controller opens the nozzle vanes 234 to reduce the engine pressure as much as possible. The back pressure acts to extend the control rod 86.

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 much lower than atmospheric pressure.

The net pressure difference across piston 270 creates a force that allows control rod 86 to be retracted, but pistons 268, 27
The surface area of 0 is designed such that the outward force acting on piston 268 is light enough 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 such that when manifold pressure approaches atmospheric pressure, such as 12 psi absolute, control rod 86 retracts into controller 84 and closes the nozzle vanes, thereby causing turbocharger 20 It is planned to increase the speed and pressure ratio obtained from the turbocharger compressor. The controller 84 is the vane 23
4 to begin closing just before atmospheric pressure is reached, since this suction manifold pressure corresponds to an open capletor throttle and therefore indicates the need for turbocharger boost. be.

Front surface 268a of piston 268, as shown in FIG. 5a.
As the manifold pressure in the pipe 88A approaches atmospheric pressure, this pressure, together with the action of the spring 276, overcomes the atmospheric pressure that cools down from the rear surface 268b of the piston 286 and the control rod 86. into the controller 84. The pressure differential across piston 270 still exerts a net force that retracts piston 27 into its cylinder, thus assisting in retraction of control rod 86 into controller 84.

As the manifold pressure increases to a value in excess of atmospheric pressure, such as 20 psi absolute, the manifold pressure acting on the rear surface 270b of the fitting 270 through the tube 88B overcomes the spring force of the spring 278. Then, the control rod 86 is moved outwardly from the controller 84 and the vane 234 is again fitted with a metal opening to prevent excessive pressure. The control rod 86 extends from the controller 84 in this pressure condition, but the piston 2
68 is held in the fully retracted position within the cylinder 265. When the piston 270 overcomes the spring force of the spring 218 and moves to the position of the piston 258, the control rod 8
6 has the ability to move freely relative to the piston 268. .

Therefore, the present controller is configured to operate the nozzle vane 23 at a manifold pressure having an absolute reading less than or equal to a predetermined value below atmospheric pressure, as determined in accordance with the schedule of the present controller and components therein.
It provides a system for opening 4. Similarly, the controller closes the vanes 234 at manifold pressures ranging from a predetermined value below atmospheric pressure to a value above atmospheric pressure, and closes the vanes 234 at manifold pressures ranging from 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, where the solenoid keeps the shaft 86 extended during light load operation of the engine and when heavy loads are required. If the situation arises, @all will be released. When the shaft is released, the splash 27
6 retracts the shaft 86 to close the nozzle and increase the speed of the turbocharger. When light pressure is delivered from the suction manifold to the face 270b of the piston 270, the shaft 86 extends to fully control the manifold pressure through the nozzle opening, as described above.

In operating a variable area turbine nozzle, a control signal, such as compressor discharge pressure, is communicated to actuator 80, which actuator appropriately controls control 886 in accordance with the signal to the actuator. extend or retract. By this, the operating lever 24 attached to the 1611 (Incorporated) bar 86
0 rotation and the outer ring 25 of the control ring 246
Two simultaneous angular rotations occur.

Further, the rotation of the outer ring 252 causes each operating lever 240 to rotate fully, and accordingly, the angle of the nozzle blade 234 is fully set.

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

Conventional uncontrolled turbochargers typically produce a boost pressure that is directly related to compressor speed, which typically corresponds to engine speed.

Therefore, the boost pressure in such conventional turbochargers is lower when the engine speed is slow than when it is high. Therefore, at low engine speeds, conventional turbochargers are not effective in improving overall 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.

A turbocharger according to the present invention can be controlled to provide boost pressure at low engine speeds. As a result, engine performance can be improved at low engine speeds, and the lag time associated with conventional turbochargers can be eliminated or reduced.

The present invention achieves this result through control of a variable turbine nozzle that maintains the speed of the turbocharger turbine and compressor at a level that provides the required boost at all engine speeds. As seen in FIGS. 6 and 7 illustrating this embodiment of the invention, the actuator 8o
Instead of the control 22L-400'E is used. The control unit 400 comprises a pulse-to-voltage converter 404 and a magnetic sensor 402 coupled to a whole embryo monitor 403. The rotation of the shaft 160 is controlled by the disk 16 attached to the shaft 160.
0'' is monitored by sensor 402.

Disk 160'' has multiple degrees of misfortune and reading.

The speed of shaft 160 is determined in a well-known manner to convert the electrical pulses produced by the motion of the disk relative to magnetic sensor 402 and then convert these pulses into total pulses.
υ is determined by the method of inputting it to the voltage converter 404.

This voltage, which is proportional to the rotational speed of disk 16σ' and shaft 160, is converted to shaft speed by monitor 403.

The monitor 403 is connected to the proportional controller 414 by suitable conductors.
Connect to. This controller 414 is coupled to a power source 416 and a nozzle area control unit 415.

This unit 415 controls the movement of the control rod 86 and rotates the lever extension 262 to change the direction of the vanes 234 in the inlet area of the turbine rotor 110.

In one embodiment of the invention, controller 414 includes a microprocessor that receives analog input from sensor 417 and converts it to a digital analog input that can be used by the microprocessor. A speed monitor 403 is provided. This microprocessor may be of the conventional type currently found in automobile ignition systems. Alternatively, the microprocessor that may be incorporated into this system may be the microprocessor model 8080 manufactured by Intel. The proportional III ral device 414 is
The control unit 415 is equipped with a total of seven amplification systems for amplifying the signals supplied from the microprocessor that controls it.

This microprocessor determines the boost pressure created by the turbine speed and therefore the turbocharger compression limit as a function of several parameters such as engine speed, throttle position, manifold suction pressure, engine torque, ambient temperature or transmission gearing. You can program it to control it. As shown in FIG.
17 measures each of the engine parameters and sends corresponding analog readings to a proportional controller 414 where the parameters are converted to digital ohms for use by the microprocessor.

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.

The proportional controller 414 shown in FIG. 6 will be explained in detail with reference to FIG. 6A. A plurality of sensors 417a-g measure all parameters related to the operation of the motor vehicle engine. Each of the sensors 417, along with the turbine speed monitor 403, generates an analog signal that is transmitted to a respective analog-to-digital converter 424a-h. The signals generated by each of these sensors are converted in a corresponding analog-to-differential converter into a series of digital words representative of the numerical value of the parameter being measured.

The dinotal signal produced by analog-to-disotal converter 424 is conveyed to a microprocessor contained within proportional controller 414. microprocessor 426
receives a digital ratio signal representative of the parameter being measured by sensor 417. The microprocessor evaluates the combination of signals received and then generates a signal corresponding to the best control position for the nozzle vane 234. The output signal generated by microprocessor 426 is sent to digital-to-analog converter 428, which generates a corresponding nozzle control signal. The output of the digital-to-analog converter 428 is connected to a driver circuit 429 that amplifies the nozzle control signal. The amplified signal is communicated to a nozzle area control unit 415 that selectively positions the nozzle vanes 234. , driver circuit 429 not only provides amplification for the nozzle control signal, but also connects proportional controller 414 to control unit 41.
5 and proportional controller 414.

Proportional controller 414 and sensor 417 in accordance with the present invention utilize technology previously developed for automotive engine control for spark timing and fuel metering. This single existing system is described in Gerald M. Walker's ``Aut
automotive electronics gets
The GreenLight J EIJCTRON
IO8, Volume 50, m20 (1977
(September 29, 2016) Pages 83-92 contain δ self-published. Thus, with the aid of the present invention, an engine designer can obtain the required boost pressure relative to one or more of these or other parameters.

Needless to say, the parameters listed above provide the best turbocharger compressor speed over the entire range of engine and turbocharger motion! -, S is merely an indication of some of the parameters a designer may wish to use to select.

The simplest arrangement is simply to maintain the turbine speed and therefore the compressor speed at a constant level throughout all engine operation. A speed is selected. Controller 414 sends a signal to actuator 418 in response to the comparison of the setpoint speed to the speed indicated by monitor 403. The shaft speed monitored by monitor 403 is selected by controller 414. 414 , the controller 414 controls the actuator 41 .
Signal full speed, vane 234 to retract rod 86 at 8.
Rotation of the turbine and closure of the turbine nozzle result in a breakout. As the turbine nozzle closes, the turbine speed increases and the shaft 160Lv rotational speed increases. If the shaft speed indicated by monitor 403 is greater than the setpoint value, a signal from controller 414 to actuator 418 causes extension of rod 86 and corresponding opening of the turbine nozzle to This will reduce the speed of the shaft 160 attached. By continuing such operations, the turbine and compressor speeds can be controlled according to predetermined setpoint values programmed into the controller 414.

Therefore, regardless of engine speed and throttle position, the blades 23
By changing the position of n7, a constant compressor and turbine speed can be maintained, and the flow rate of the exhaust gas hitting the turbine rotor 170 can be completely controlled.

By using the turbocharger shown in FIG. 2, a relatively high compressor speed can be maintained with respect to engine speed.

This is possible because of the ability to completely change the surface profile of the turbine nozzle. Therefore, the engine fully accelerates! There is no ramp-up time for the boost pressure, since there is no need to increase the compressor speed from a relatively low speed to a high value that would otherwise be kept 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. The advantage gained by the variable area turbine nozzle with the turbocharger shown in FIG.
The use of a compressor that can be operated over the L range is significantly amplified. In other words, conventional turbocharts have been plagued by stall lines that completely prohibit compressor operation over a wide flow range.
The use of variable area turbine nozzles would not provide the same benefits as in the turbocharger shown in FIG.

Accordingly, the turbo gear shown in FIG.
It is a turbocharger with a large turbine nozzle area. Additionally, the control ring and actuating lever are located in unused air ear space between the turbocharger turbine and compressor forming a very compact unit. Furthermore, the special location of the linkage isolates 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 economically cheap to manufacture and operate, and reliable.

Furthermore, the use of ball bearing supports and the elimination of corresponding thrust bearings and disc-type bearings provides a more controlled or stable turbine-compressor rotating system, which allows both the turbine and the compressor to be connected.
Only a narrow gap is allowed between them and surrounding structures. For example, the height of compressor blades for turbochargers 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 a 0.005 inch clearance, or 2.5% of the total blade height. On the other hand, if a straight or disc type bearing were used, a significantly larger clearance (typically about 0.015 inch) would be required. As a result, the turbocharger shown in FIG. 2 is particularly suited to turbocharger construction for small internal combustion engines where low blade heights require narrow gaps between the blades and the surrounding grooves. ing.

Due to the use of ball bearings and the resulting reduction in friction loss in the case of soft disk bearings, a considerable portion of the turbine horsepower is not lost due to friction. For the first time, it is possible to effectively use a turbocharging shaft that generates less turbine horsepower.Considering that journal bearings and disk type bearings account for 4 to 5 horsepower losses due to friction, the friction losses due to the arrangement of the present invention are 0
．． It would be on the order of 1 to 0.4 horsepower.

FIG. 8 shows the compressor rotor 172 and the turbine rotor 11.
The retaining nut 180 and the compressor-turbine shaft 160 are shown off the ground. Turbine rotor 170 is attached to one end of shaft 160 by welding or other suitable permanent attachment means.

The shaft 160 has an enlarged bearing surface 160a and a stepped portion 18 reaching the small shaft diameter portion 160b before being added to the rotor 170.
8. As mentioned above, the races 206 and 210 are directly formed on @160.

During assembly, shaft 160 is inserted through holes in compressor back wall 100 and turbine back wall 102. Shim 186 to axis 1
It is positioned on @160 so as to be associated with the stepped portion 188 of 60. Compressor rotor 172 is associated with shaft 160 and rests on detail 160b. Then the retaining nut 180
is pushed into the hole 182 of the entire rotor 172 on the shaft portion 160b and associated with the bottom wall 184 of the hole 182. Nut 180 consists of a smooth inner bore 180at a provision sleeve. The hole that goes through 180 (oak) is shaft 1
The part 160b of 60 is fully formed with an interference fit. This interference fit is on the order of 0.001 inch in one embodiment of the invention.

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, this 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 way, the cost and problems associated with forming threads on heat-treated shafts are eliminated; furthermore, since the shaft is considerably hardened, the retaining ring 180 can be pressed against the shaft or It can be pulled out without damaging the surface.

As shown in FIG. 8, nut 180 is threaded about its outer surface. Hole 182 is of a diameter just wide enough to provide clearance 420 (FIG. 9) between the nut 180 and the threads on the outer surface thereof. This gap 420Ta makes it possible to insert a suitable internally threaded tool for pulling out the compressor rotor from the shaft in order to remove it.

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 can be mounted in place with high compressive loads within the compressor rotor or tension fittings within the compressor-turbine shaft. It does not have the ability to actually occur. Therefore, the first
In figure 0, hold Nara! -180 and the compressor rotor 172 is a conical or Bellevueille spring 422t- inserted.

The components in Figure 10 are the same as or equivalent to the parts in the embodiments of Figures 2 and 8 and 9 and are therefore designated by the same numbers marked (') in the latter figures. Parts corresponding to or similar to those in the examples are shown. In FIG. 10, the shaft 160' is the compressor rotor 1T.
It passes through 2'. The retaining nut 180' is attached to the shaft 160
Bellevue spring 422 mounted on shaft 160' at the end of
and oak) 180' and hole 182' of rotor 172'
The walls 184' are associated with each other. Bellevue spring 42
2 is when attaching the holding nut (180') to the shaft 160'.
Originally, it was in a state of pressure. The connection of the oak 180' to the shaft 160' is such that the connection between the oak 180' and the rotor 1γ2' is sufficient to overcome any expansion force created by the Bellevue spring 422. Instead, the spring 4 between the rotor 172' and the oak 180'
The compressive force of 22 induces an axial curve l in rotor 172'. Therefore, when spring 422 is in its normal position, shaft 1
60' or rotor 112' (Nara)
The coupling force V between rotor 180' and rotor 172' becomes zero.

FIG. 11 shows the compressor housing 24, the compressor front wall insert 126, and the respective molds τ used to die cast each of these two pieces. Traditionally, many turbocharger compressor housings are sand cast to provide a circumferential chamber through which the compressed air is directed into the intake manifold of the engine with which the sheet turbocharger is used. , which 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 using sand mold νJ construction requires higher column costs than making it by die casting, and creates a structure with a low sheathing degree that requires a rough wall structure within the circumferential direction. It leads to However, in the past when turbocharger compressor housings were made by die casting, it was rarely possible to make the front diffuser wall as was possible with 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, it is possible to provide a two-component die-cast compressor housing that has advantages in assembly that are conventionally available only by sand casting. The compressor housing 24 is formed using an outer mold 29o1, an inner mold 292, a core mold 294, and a cap mold 296. Inner mold 292 is given a raised profile to form circumferential chamber 44 . In addition, this inner mold 292 has a chipped portion 300.
The artificial hole 40, the blade 62, and the bearing support cylinder 60 are formed by combining with the protruding portion of the corresponding outer mold 290. Core mold 294 and cap mold 2
96 is used to form the exhaust port 52 of the pressure m iso.

As is clear from FIG. 11, the molds 290, 292°2
94.296 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. The molds 320, 322 work together to produce the compressor front wall insert 126. The molds 320, 322 have abutting surfaces that interlock to define the separation line 324 between the discs 1 and 32.
Create an insert 126 with 30 outer edges. 11th
Throat 128 includes a notch 32, as clearly shown.
6 is provided so that the insert 126 receives the post 62 when installed within the housing 24.

Figure 12 and cutting drawings 13, 14 and 15 show the blade 62.
It shows the positioning and shape of. 16th to 15th
In the figure, each strut 62 has a thicker intermediate center cut 3
It has a leading edge 330 with #34 and a trailing edge #332.

In each case, the thickest center cut is as indicated by line 336 and 6 is connected to mold 290 used to form housing 24.
, 292. That is, the struts 62 are formed by die casting using molds 290, 292 to create the desired air-oil shape of the leading and trailing edges separated by a thicker center cut therebetween. This shape significantly facilitates the flow of air into the compressor inlet area, as shown in Figures 13-14.15.
The shape of the present invention can be cast using a well-known die casting technique.

In order to make the airfoil shape of the strut 62 taper from the thicker center cut to the thinner leading and trailing edges, the die-casting mold 4 has inlets from both the leading and trailing ends. It must be inserted into hole 40. That is, as shown in FIG.
The first inner wall 110 (FIG. 2) is formed by the mold 290, whereas the first inner wall 110 (FIG. 2) is formed by the protrusion 300 of the mold 292. The inner cylinder walls 110, 112 each expand outwardly to allow the die casting mold to be removed after the piece is formed.

Thus, the formation of the required geometry of the struts 62 requires a wall 112 with a diverging diameter within the entrance hole 40 of the die casting die receptacle. However, it is important that the inlet hole 40 has a diameter that decreases continuously from the inlet toward the compressor rotor 112. This means that wall insert 1
This is done by using 26. wall insert 12
6 is also formed by die casting, and the inner wall portion 112 of the inlet hole 40
It has an outer surface 132 (FIG. i2) that interlocks with the diameter of the flared surface corresponding to the diameter of the flared surface.

The inner wall surface 131 of the throat portion 128 corresponds to the tapered diameter extension of the first inner wall portion 110 of the inlet hole 40 .
It is formed to have a diameter that tapers in the direction of 30.

That is, when wall insert 126 is combined with inlet hole 40, it provides a diameter that tapers continuously inwardly from the inlet of hole 40 toward compressor rotor 112 (FIG. 2).

In this way, a tunibead structure forming a compressor housing for a turbocharger is obtained, both of which can be made 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 a thicker middle that decreases toward a thinner leading edge and a thinner trailing edge. It has blades with cutting parts. Furthermore, the housing forms an artificial nozzle that is successively smaller from the entrance of the inlet hole toward the compressor rotor. Additionally, the compressor housing defines a front diffuser wall that completes a circumferential chamber through which the 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 a more restricted airflow into the turbocharger, while wall insert 126b forms an even larger compressor rotor and provides even more airflow into the turbocharger. These structural changes allow different wall inserts 126 to be used with standard compressor housings 2.
It will be seen that this can be done simply by attaching it to 4. That is, among several possible shapes for the wall insert 126, each with a corresponding compressor wheel, using a variety of different turbocharging layers z1 standard compressor 24, each with different flow performance. Create by selecting one shape. 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.

Accordingly, it is 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 compressor housing. Such a two-part construction allows the valleys of these two parts to be die-cast using standard die-casting techniques. The 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. made as a gift. 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 artificial hole. The vanes are formed with leading and trailing edges separated by a thicker center cut. This is done by using a mold which is provided with a separation line in the approximately thicker transverse section to enable die casting of the compressor housing.

Die-casting the compressor housing in two components is used in a turbocharger in which the compressor rotor and turbine rotor project to and are supported on one side of a shaft that rotates on a ball bearing arrangement. However, it will be obvious to those skilled in the art that the present invention 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 compression housing as described above can be used as is by simply modifying the bearing support cylinder.

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 the compressor speed and flow are wide, The ability to operate without experiencing the surging lines normally found in conventional turbochargers.

This extremely significant and important feature will be explained with reference to a performance graph known as the compressor map shown in Figure 17.18. Figure 17 is a typical compressor map for a conventional turbocharger;
The figure is a compressor map for the improved turbocharger. These compressor maps are
Figure 2 shows the relationship between the cubic volume of air flow per minute and the ratio of discharge pressure to inlet pressure for turbocharger compression.

FIG. 17 shows a constant speed line 430 and an isentropic efficiency line 432 for the compressor.

Surging line 434 defines the operating point, to the left of which the turbocharger compressor cannot operate to produce -like air output. Turbocharger components can be modified to allow stable operation at lower turbocharger speeds, i.e., lower engine speeds, if the surging line is moved toward the vertical axis.
This shift in the operating characteristics of the turbocharger reduces the high speed capability of the unit. Accordingly, turbochargers have traditionally been downgraded to provide increased performance only at higher engine speeds.

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

As you can see from the compressor map in Figure 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 φ map of FIG.

By looking at Figures 17 and 18, the width of the operating range of the conventional turbocharger can be compared with the improved turbocharger. According to Figure 17, at a compressor pressure ratio of 1.9, the air flow range is 65 at an efficiency of 60%.
It varies from a high of 0 cubic vacuum to as low as 200 cubic vacuum per minute at the surging line. Therefore the ratio of high flow to low flow (
350÷200) becomes 1.75. According to FIG. 18, at a pressure ratio of 1.9, the improved turbocharger operates from a flow rate of 275 cubic meters at 60% efficiency to 60 cubic meters per minute without encountering a surge line. The ratio of the high and low flow rates (275÷60) is 4.58, which is an increase of more than 2.5 times the increase in conventional turbochargers.

These achievements are of great significance in that turbochargers can be effectively used to improve their performance at both low and high engine speeds. Such advancements, when combined with the features of the present invention that provide constant compressor speed for varying engine speeds or constant dest at varying engine speeds, can provide even higher desired pressure ratios at lower engine speeds. It is also significant in that it is obtained without the surging or turbulence problems typically experienced with lower compressor flows in conventional turbochargers.

Such significant advances provided by the improved turbocharger are primarily due to the improved radius α1f,
In combination with type compressor technology, such as an aft-curved blade arrangement, the pressure S (I) should be attributable to the air inlet shape. As already mentioned, the bearing support surface 60 It is supported in the hole 4υ by a plurality of struts 62 extending from the inner wall of the hole to the bearing support tube 60. In one embodiment, struts 61111 extend from the inner wall of the inlet hole to the bearing support tube 60, and this support tube is connected to the compressor. located concentrically within the air inlet of the inlet, so that the inlet is divided into one or more inlet channels, which channels are also proportionally long in the flow direction compared to the width dimension; Each strut has a profile with leading and trailing edges that are narrower than the intermediate section of the strut.As also shown in FIG. It has a diameter that tapers to .

This structure stabilizes the backflow conditions normally encountered at low compressor flows. By stabilizing backflow conditions, the compressor can continue to operate to produce steady airflow at much 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. The turbocharger of FIG. 19 is the same as the turbocharger illustrated in FIGS. 1-16 except that an axial membrane compressor rotor 450 is mounted within the compressor inlet. 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 connected to the centrifugal compressor rotor 1.
72 is modified to include a larger diameter inlet 452 with an inner wall 454 defining an inlet passage 456- to 72.

A plurality of stators 458 extend from the inner wall 454 of the inlet 452 and support a bearing support 460. Axial membrane compressor rotor 45
0 is each bearing assembly! 7162.164 is attached to the end of the shaft 160 on the opposite side of the centrifugal flow compressor rotor 172 and is retained by a suitable nut 462. Compressor rotor 450 includes a boss 464 and a plurality of struts 466 attached between boss 464 and struts 468. boss 4
64fC1 is associated with step 470 of shaft 160 and located between step 470 and nut 462.

As is clear from FIG. 19, the air entering the inlet 452 is
It is compressed by the axial membrane compressor and conveyed to the centrifugal compressor rotor 112 via the flow path 456. The use of a two-stage compressor is important in that the use of a two-stage compression chamber provides a wider operating range than that available from a single-stage compressor.

For example, when using a two-stage compressor, the first stage is used to obtain a pressure ratio of inlet pressure to discharge pressure of 1.5 to 1.6, while the second stage is used to obtain a pressure ratio of 6.0 to 6.5. Used to obtain ratios.

These combined pressure ratios effectively provide pressure ratios on the order of 4.5 to 5.6. Such pressure ratios can be obtained with a single stage, but the higher compressor speeds required to produce such pressure ratios necessitate the use of higher quality and more expensive compressor and turbine components. Moreover, the flow range is severely restricted. With the construction of the present invention, a two-stage unit provides better pressure ratios without this additional cost or disadvantage. Furthermore, the present invention provides that mounting the axial membrane compressor on the opposite side of the bearing support from the centrifugal compressor and the valley rotor of the turbine does not result in any particular adverse effects of balancing the system.

FIGS. 20, 21, 22 and 26 show a new method of attaching the turbocharger illustrated in FIGS. 1 to 16 to an internal combustion engine. Figures 20 and 21 are v
- A hard sectional view and a plan view of the turbocharger 2o attached to the 66 gate 444 are shown, respectively. Of course, this same application can be applied to V-type engines having any number of cylinders. Engine 444 represents a conventional V-shaped engine having 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 #445 and the intake valve 449 of the left cylinder group 446. The turbocharger 20 is mounted in an orientation that is not conventionally possible in conventional turbochargers, with the rotational axes of the turbine and compressor in a rigid position. The turbocharger 20 has a corresponding structure that does not require a continuous flow of lubricant to the core material and the bearings, and is designed to deliver oil to the bearings, so that the axes of rotation of the compressor and turbine are aligned vertically or vertically. The turbocharger can be oriented at any angle between horizontal lines. On the other hand, conventional turbochargers, which normally require a flow of lubricant to and from the bearings, have been successfully operated in a vertical orientation, as exemplified by the configurations of Figures 20 and 21. Not yet.

As shown in FIGS. 20 and 21, the engine exhaust gas is guided to the turbocharger 2o by a manifold 451. In this case, the exhaust gas passes directly from each exhaust valve 448 to the turbine rotor 170 of the turbocharger 2o. The exhaust gas from the turbocharger compressor passes through each cylinder of the engine 444 via a pipe 453 to the intake valve 449 via a manifold I'447.
sent to.

During operation, air enters the turbocharger 20 through the inlet 50.
The air is sucked in by the compressor rotor 112, passes through the compression wrinkle pipe 453, enters the manifold 441, and is discharged into each cylinder of the left and right groups of the engine 444. Exhaust from the engine 444 is directed to the turbine rotor 1 by a manifold 451.
70 at mutually spaced positions around it,
It is then discharged through the exhaust pipe 455. According to this interval,
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 aligned with 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. The 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 highly complex manifolds to direct exhaust gas from the exhaust port to the turbine rotor. The turbocharger 20 further includes an exhaust manifold r that directs exhaust gas from each exhaust port from the engine to approximately equally spaced locations around the turbine rotor and to different locations around the periphery of the turbine rotor. Send exhaust. Because the turbocharger 20 can be vertically oriented, the turbocharger 20 can also be located close to the engine between each group of cylinders. Additionally, 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 26 show the case where the turbocharger 20 installed in a vertical position is used in a parallel four-cylinder engine. Of course, the turbocharger 20 can also be used in parallel engines with more or less than four cylinders.

In FIGS. 22 and 26, the turbocharger 20
The engine 480 is operated by an exhaust manifold 482 that communicates with the turbocharger turbine between the exhaust valves of each engine cylinder.
It is connected to. Air is drawn into the turbocharger 20 at an inlet 50 and the turbocharger compressor rotor 17
2, and is released into the intake valve of each engine cylinder through the manifold 1484. As seen in FIG. 26, the exhaust manifold 482 is configured to direct the gases that drive the turbocharger turbine to the turbine at spaced locations around the turbine.

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 between two cylinder groups of a V-type engine.

Therefore, a turbocharger is obtained that 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 compressor and turbine shafts in the turbocharger 20 are designed to lubricate the rotating shafts. Because turbocharger 20 can be oriented in this manner, it is ideally suited for use in conventional engines of all types and constructions.

The turbochargers described in FIGS. 1 to 16 are shown in FIGS. 24, 25, 26, 27, and 2.
It can be easily converted into an effective turbojet engine or turboblower engine as shown in FIG. Although these embodiments illustrate the vertical cross section of only the upper half, it goes without saying that the lower half is almost the same as the upper half shown.

The turbojet 500 is illustrated in FIG. Turbojet 500 includes an outer housing 502 consisting of a tubular inlet nozzle 504, a main housing 506, and a combustion chamber rear wall plate 508.

The rear wall plate 508 includes integral reinforcing members 510 and bolts 51.
2, it is attached to the main housing 506. The turbine exhaust nozzle 514 is supported within the main housing 506 by attaching it to the end of the rear wall plate 508 opposite the main thread 506 by bolts 515 .

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 tubular inlet nozzle 504 . This structure is similar to that illustrated in FIGS. 1 to 16 with respect to the bearing support body 60 and its attachment within the inlet 40 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 main housing 506 and are attached to each other by suitable threaded pieces 527. In the embodiment illustrated in FIG. 24, a fixed stationary vane 528 is mounted between the main housing 506 and the compressor back wall 524 by a suitable threaded piece 529. combustion liner 530
is attached between the turbine rear wall 526 and a laterally extending portion of the turbine exhaust nozzle 514. Combustion liner 530 is of conventional construction with a circumferential chamber defining a plurality of ports. The liner 530 has a port 532 for primary air combustion and a port 5 for dilution air according to known practices.
It is equipped with 34. Fuel is supplied to the combustion chamber by fuel injection nozzles 536 that extend through the combustion chamber rear wall plate 508 and the upstream wall 538 of the combustion liner 530. igniter 540
The upstream wall 538 and the combustion chamber rear wall plate 508 are housed within the combustion liner 530. ``The combustion liner 530 includes a turbine inlet area 544 formed between the turbine rear wall 526 and a laterally extending portion of the turbine exhaust nozzle 514.
It is provided with an opening 542 that leads to.

Turbine inlet zone controller 546 is used in turbojet 500 illustrated in FIG. 24 and is the same as turbine inlet zone controller 228 illustrated and described for the turbocharger of FIGS. 1-16. In particular, the turbine inlet zone controller 546 includes movable nozzle vanes 548 located circumferentially about the nozzle zone and pivotable to vary the velocity and angle of the flow of combustion gases through the nozzle inlet zone 544. . The movable nozzle vane 548 includes trunnions 550 and 552 extending from opposite sides thereof. The trunnion 550 is the turbine exhaust nozzle 5
14 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
Turbocharger 20 illustrated in FIGS. 1 to 16
As discussed above, the inner race of the ball bearing assembly 556 is secured to the compressor rear wall 524 in conjunction with the outer race of the ball bearing assembly 556. The action of the drive lever 554 is also the same as that described for the turbocharger 20 of FIGS. 1-16 to control the angular orientation of the movable nozzle vanes 548 as desired.

A bearing support cylinder 516 is supported concentrically within the inlet nozzle 504 by struts 518 as described above. The compressor rotor 570 and the turbine rotor 572 are mounted on two ball bearings 5.
76.578 is supported on a shaft 574 mounted for rotation within bearing support cylinder 516. A turbine rotor 572 is attached to one end of the shaft 574, and a centrifugal compressed steel rotor 570 is attached between the turbine rotor 572 and each ball bearing assembly 576,578. axis 57
4 passes through the turbine aft wall 526 and the compression I aft wall 524. In this case, a labyrinth seal 580 formed around the shaft 514 provides a seal between the compressor 1o-p 570 and the turbine rotor 572.

The turbine rotor 572 is fixed to the υ shaft 574 by welding or the like. The compressor rotor 570 is also held on the shaft 574 by a holding nut 582. Nut 582 is the nut 18 described for the embodiment illustrated in FIGS. 1-16.
It has the same structure as 0.

Compressor rotor 570 is connected to stage 58 of nut 582 and shaft 574.
It is located between 6 and 6.

As shown in FIG. 24, the support ring 590 is attached within the end of the cylindrical body 516 adjacent to the compressor rotor 570, and the cylindrical body 5 is supported by the retaining ring 592 attached to the cylindrical body 516.
It should not move within 16. Ball bearing assembly 5
The outer race 594 of 7B is formed into a support ring of 590 yen. Inner race 596 is integrally formed with shaft 574. Balls 598 are housed between inner race 596 and outer race 594 to form ball bearing assembly 578.

The bearing support cylinder 516 supports the inflow nozzle 504 by a plurality of struts 518 extending from the inner wall surface 520 of the inflow nozzle 504.
It is installed in 04. This structure is based on the bearing support cylindrical body 6
0 and its accommodation in the inlet of the turbocharger 20 are similar to the structures illustrated in FIGS. 1 to 16. 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 I/Ujifugu 506 and are attached to each other by threaded pieces 521.

The ball bearing assembly 516 is slidable within the cylindrical body 516 and the same race 600 integrally formed on the shaft 574, and the ball 60 is slidable within the cylindrical body 516.
6, and an outer ring 602 having an outer race 604 formed therein to receive the outer ring 602.

A compression spring 610 is associated between the outer ring 602 and a retaining ring 612 secured within the cylindrical body 516 to urge the outer ring 602 outwardly to force the balls 606, 598 in each ball bearing assembly 576, 578, respectively. By fixing the position, the shaft 57
Fix position 4.

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

In operation of the turbojet 500 illustrated in FIG. 24, air enters the tubular air inlet at its port 620 and follows the air path indicated by arrow 622. This air is compressed by the compressor rotor 5γ0 and is then discharged radially along the air path indicated by arrow 624. This compressed air is pumped into the combustion chamber section 626 and into the combustion liner 530 via channels 532,534 each day. The air is
It mixes with the fuel injected into combustion liner 530 through nozzle 536 . In this way, the stoichiometric mixture is returned. This mixture is ignited within combustion liner 530 by igniter 540 . Cooling air is directed into the combustion liner 530 through a port 534 to cool the combustion exhaust gas before it enters its turbine/inlet area 544. These gases are associated with the turbine rotor 572 and the shaft 57 is connected to the rotor 572.
4 and the attached compressor rotor 570.

Next, the fuel a and f gas are discharged to the outside air through the nozzle 514 to generate thrust by the 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 turbine and compressor rotor shafts are located away from the rotating bearing turbine rotor 572 and from the extremely hot combustion gases that are delivered from the combustion liner 530 to the rotor 572.

Furthermore, the practical problems caused by the extremely high temperatures normally experienced by such bearing structures are usually completely eliminated by the unique bearing arrangement described above. The bearing uses the fuel used to drive the turbine.
Limited lubrication is sufficient for effective operation of the turbocharger since the compressor rotor and compressor rear wall and turbine rear wall are virtually insulated from the extremely high temperatures associated with the burning gases. . this. The need for a constant flow of lubrication to the mating bearings, as well as the many parts necessary to provide such lubrication that must be driven by the rotating shaft, are eliminated.

FIG. 25 shows a turbo blower 650. turbo blower 6
The structure of the turbojet 500 is the same as that of the turbojet 500 illustrated in FIG. 24, except that a bypass air passage is provided in addition to the primary air passage, and a nine-stage shaft compressor is provided upstream of the centrifugal compressor. Almost the same.

As shown in FIG. 25, turbo blower 650 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 secured to the combustion chamber rear wall 656 by bolts 662.

The compressor back wall 664 and the turbine back wall 666 are located within the main shaft 654 . And burning litiro 6
8 is installed in the combustion chamber 670, and the opening 612 opens between the turbine rear wall 666 and a portion 674 extending in the direction of the turbine exhaust nozzle 660. Combustion liner 668 has a plurality of fliil ports 676 formed therein. The fuel injection unit 618 penetrates the rear wall 656 of the fuel g8 chamber and
It is stored inside. 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 mounted 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. Also, a plurality of stationary blades 688
is attached to the inlet nozzle 652 for cooperation with an axial membrane compressor, as will be described in more detail below.

Shaft 690 is mounted within bearing assembly 6 within bearing support cylinder 684.
It is supported by 92,694. Each bearing assembly 69
2, 694 is Tahozoera in Figure 24) 500 and 1st
This is the same as described with respect to the turbocharger illustrated in FIGS. Inner wall 69 of inflow nozzle 652
An oil chamber 696 is formed between the bearing support cylinder 684 and the bearing support cylinder 684.

A suitable ring 100 with 0-shaped rings 702 and 704 is chamber 696.
It is located in the mouth of the patient and is held by a ring 706. Each of the lamp cores 708, 710 communicates through an oil chamber 696 to send oil to ramps 712, 714, respectively. Slip ml of 692 and 694.

Medium flow compressor rotor 116 and turbine rotor 718 are attached to shaft 690 . Figures 1 to 16 and 2
The same turbine inlet zone controller 720 as described for each embodiment in FIG. 4 cooperates to control the velocity and angle of the combustion gases to the turbine rotor 718.

Bypass flow path 722 is formed between bypass inner wall 724 and bypass outer wall 726. Bypass outer wall 726 is attached to bypass inner wall 724 by a plurality of fixed stationary vanes 728 positioned between the two walls. The main housing 654 is positioned by a plurality of columns 730.

The axial flow membrane compressor rotor 732 has the rotor 1 attached to the retaining ring 736.
32 is attached to the shaft 690 by a suitable nut 734 that abuts the shaft 690. The rotor 732 includes a boss 138 and a plurality of blades 740 extending radially from the boss 138.

The plurality of cylindrical armatures 142 are connected to the boss 1 of the rotor 132.
A fixed field winding 744 embedded in and mounted adjacent to the bypass wall 726 to a post 746 cooperates with it to produce the current.

The operation of the turbo blower 6500 illustrated in FIG. 25 is similar to the operation of the turbo blower 6500 except that the air entering the inlet of the turbo blower 650 by the axial stage compression rotor 732 is divided into a bypass flow shown by arrow 747 and a primary flow shown by arrow 748. The operation is almost the same as that of the turbojet 500 illustrated in FIG. Bypass air is compressed and discharged through bypass nozzle 749 to generate thrust from turbo blower 650. The primary air is directed along the path indicated by arrow 148 and passes through the axial stage compressor rotor 732 and the centrifugal compressor rotor 7 before mixing with the fuel in the combustion liner 668 and igniting within the liner 668.
16. The ignited gas flows past the turbine inlet zone controller 720 toward the turbine rotor 718 and along with the axial film compressor rotor 73 .
2 and centrifugal flow compressor rotor 716. Exhaust gas from the turbine rotor 118 is passed through the turbine exhaust nozzle 66
0 and provides additional thrust to the turbo blower 650.

The turbine inlet area control system 720 is shown in FIGS. 1-16.
This is the same as described with reference to the illustrated turbocharger 2o and the turbojet 500 shown in FIG.

Controller 720 operates to control both the velocity and angle of the combustion gases directed from combustion liner 668 to turbine rotor 718 . Of course, the monitoring devices described in connection with FIGS. 6 and 7 with respect to the turbocharger 2o apply directly to the turbojet 500 and the turbo blower 650 of FIGS. 24 and 25, respectively.

i.e. monitoring the speed of the turbine and cooperating compressor;
The turbine nozzle inlet area is varied according to this speed or other engine parameters as desired.

During rotation of the rotor 732 of the rotor 732, the armature γ42 rotates relative to the winding 744 and generates electricity.

This electricity powers loincloth engine or aircraft equipment as desired, resulting in an extremely economical and safe energy package for this equipment.

26 and 27 show a turbochet 150 as a modification of the p-g jet 500 illustrated in FIG. 24. The turbojet 750 has a main housing 75
The main housing 15 includes an inlet cylinder 752 joined to the main housing 15.
4, a rear plate 756 is attached with multiple bolts 758. Turbine exhaust nozzle 760 is secured to a suitable bolt 162.
It is attached to the end of the rear plate 756 on the opposite side of the main housing 754. Turbine exhaust nozzle 760 includes a laterally extending portion 166 formed to join exhaust segment 164 and main housing γ54.

Bearing support segment 768 supports inlet cylinder 152 by a plurality of fixed stationary vanes 110.

Shaft 772 is supported within cylinder 768 by bearing recesses 74, 776. Each bearing assembly 774.77
6 is a bearing 576.5 illustrated in the embodiment of FIG.
It has the same structure and function as 78. Turbine rotor 7T
8 is formed integrally with the shaft 772. The compressor rotor 780 includes a turbine rotor 1γ8 and each bearing γ14. Tγ
6 and is attached to the shaft 7T2.

Compressor back wall 782 and turbine back wall 784 are mounted within main housing 154 . A plurality of stationary vanes 786 are attached between the main housing 754 and the compressor rear wall 782 by threaded pieces 788 .

A suitable igniter γ 89 is located within a laterally extending portion 766 of the turbine nozzle γ 60 between the segment 66 and the turbine rear wall 78.
It is installed in the room formed between 4 parts.

Laterally extending portion 766 and turbine rear wall γ, respectively
A turbine nozzle inlet zone control system 790 is provided with a movable vane 792 rotatably supported within each trunnion 794, 796. Turbine nozzle inlet zone 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 18
A labyrinth seal 800 is formed within 4 to effectively seal between the turbine section and the compressor section. Compressor rotor 18
0 is attached to shaft 7γ2 by a nut 802 that abuts rotor 780 against step 804 of shaft 772.

The axial compressor 806 rotates with the shaft 772 and the stationary blades 7
In cooperation with γ0, the inflow cylinder 152 compresses the air directed into the compressor inlet 808 formed by υ. 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 hole 812 in the 4111 line direction. 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.

An annular groove 82 extending through shaft 772 and forming a plurality of radial ports 818 in bore 812 and compressor rotor 780 .
0 and communicate with each other. A plurality of radial ports 822 communicate from groove 820 through compressor rotor 780. 0
Suitable sealing members, such as half rings 826, 828,
rotor 780 and shaft 7 mounted in an annular groove in 80;
72 to prevent fuel leakage at the connection between the two.

Turbojet T50 illustrated in FIGS. 26 and 27
During operation, the fuel is supplied to the shaft 7120 and the fuel supply pipe 8 to the hole 812.
14. Between the hole 812 and the air flow path indicated by arrow 810 (FIG. 26) is a port 818 and a groove 82.
0 and the port 822 to form a fuel supply path. This fuel flow path is indicated by arrow 838. When the radial compressor rotor 780 rotates, 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, air is drawn into the inlet 808 by the axial membrane compressor 806 and is sent out in a compressed state through the air path. This air is further compressed by radial compressor rotor 780 and mixed with fuel downstream of rotor 780.

The fuel becomes highly tinned by the action of the compressor rotor blades on the fuel as it exits the port 822 through the radial compressor rotor 780 . This atomized fuel mixes with the compressed air fed into the turbojet 750, is directed along the air path, and is ignited by the igniter 836 upstream of the turbine rotor γT8. The exhaust resulting from combustion passes the blades of turbine rotor 118 and drives the rotor in a conventional manner. These exhaust gases are then discharged through nozzle γ60 to generate thrust from 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 components previously required in a conventional turbojet. The unique fuel supply system illustrated in FIGS. 25 and 26 eliminates the need for all hardware associated with a fuel pump. Instead of an ordinary fuel pump, the fuel is injected into the air stream compressed by this rotor through a hole in the compressor rotor shaft and through the compressor rotor itself. It uses a fuel system that automatically releases . The apparatus shown in FIGS. 25 and 26 has a two-stage compression structure in which an axial membrane compressor is added upstream of a radial compressor, that is, a centrifugal compressor. This construction provides additional strength to the system by providing an axial stage compressor on the side of the bearing opposite to the centrifugal compressor.

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 oil equipment. 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 turbochet 750 illustrated in FIG.

The structure of the turbo air blower a900 is almost the same as the turbojet 750 illustrated in FIG. 27, except that a bypass desired air flow path is provided on the female side of the primary air flow path. Turbo blower 900 includes an inlet body 902 joined to a body 904 and a rear wall 906 attached to main housing 904 by ports 908 . A turbine exhaust nozzle 910 is supported from the rear wall 906 by bolts 912 and the nozzle 910
A laterally extending portion 913 of is associated with the main housing 904 . The bearing support cylindrical body 914 has a plurality of support columns 91
6 is supported from the inflow cylinder 902. A compressor back wall 918 and a turbine back wall 920 are supported within the main housing 904 . The same turbine inlet zone control 921 as described for the control 190 of the turbojet 750 of FIG. 26 is provided. Shaft 922 is supported for rotation within bearing support cylinder 914 by respective bearing assemblies 923, 924. Each bearing assembly 923, 92
4 is the same as the bearing described for the turbojet 750 illustrated in FIG. A turbine rotor 926 is attached to one end of shaft 922. Also, compressor rotor 92
8 is a turbine rotor 926 and each bearing assembly 923 .
It is installed between 924 and 924.

In FIG. 28, the axial compressor rotor 930 has a shaft 92
2 and directs air into the primary flow path within the inflow cylinder 902 and the secondary flow path formed between the secondary flow outer wall 932 and the secondary flow inner wall 934. Secondary flow outer wall 932 is supported by a plurality of stationary vanes 936 from inner wall 934 . Air directed through the secondary flow path as shown by arrow 938 is directed to compressor rotor 928 and stationary vanes 93.
6 produces a thrust from the letterbo jet which is compressed and discharged by the discharge nozzle 940. 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. The igniter 946 ignites the light. 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 910 is housed within the compressor inlet nozzle. As shown in FIG. 29, a generator 970 includes an armature 97 attached to the end of a compressor and turbine 974
2. As shown in FIG. 29, shaft 974 has a bearing 916 and a second bearing (see FIG. (not shown) is a good support. Field winding 918 is mounted within bearing support housing 980. No Ujing 98
0 is mounted concentrically within the compressor inlet nozzle 982 by vanes 984 in the same manner as described above for the turbocharger, turbojet, and turboblower embodiments. A cage 986 is attached to the end of the bearing support housing 980.

When the generator 910 is in operation, the armature 972 is connected to the field winding 9
18 with shaft 914 to produce an electrical current in a well known manner. The current thus produced is conductive!
(not shown) directs all of the power within the device from field winding 978 to any location where it is needed. That is, when the turbocharger uses the device shown in FIG. 29, the generator is used to supply current to any part of the vehicle or other device that requires electrical power. When the structure described in Section 29 is applied to the turbojet of FIGS. 24 to 28, electrical energy can be directed to the induction device or other components requiring 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 clean 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 device by directly supplying fuel from the compressor rotor shaft to the compressor rotor in the column. In this case, the fuel is released by centrifugal force, and the fuel pump and attached 7
・-Doware is not provided.

A bypass flow and a primary flow are created in the turbocharger forming the turbo blower.

In this case, a trickle stage compression chamber is also provided upstream on the opposite side of the bearing from the centrifugal compressor.

Also the pressure [4 and q! at the end of the shaft of the turbine! A generator is provided with an armature bias mounted on the opposite side of the rmll bearing compressor and turbine, and 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 the drawing]

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. 6 is a perspective view of a turbocharger equipped with a bearing support structure of the present invention. FIG. 4 is an enlarged vertical sectional view of the embodiment, and FIG. 4 is an enlarged plan view taken along line 4-4 of FIG.
5A is a partially cutaway end view of 5A-5A when viewed in the direction of the arrow in the figure.
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 diagram for explaining the control device of FIG. 6. Block configuration diagram, Fig. 7 is a vertical sectional view taken along line 7-7 in Fig. 6, Fig. 8 is a block diagram showing the turbine inlet,
An exploded perspective view showing the compressor rear wall, turbine rear wall, compressor bushing, compressor rotor, and retaining sleeve; FIG. 9 is a sectional view taken along line 9-9 in FIG. 2; FIG. 10 is a cross-sectional view showing the retaining sleeve and Figure 11 is an exploded perspective view of the two components of the housing for the compressor, with the housing separated from the mold used to make the components. Figure 12 is a front view of the turbocharger shown in Figure 1, excluding the inlet, and Figure 16 is a front view of the turbocharger shown in Figure 1.
14 is a sectional view taken along line 13-13 in Fig. 2, Fig. 15 is a sectional view taken along line 14-14 in Fig. 12, and Fig. 15 is a sectional view taken along line 13-13 in Fig. 12.
16 is a diagram showing the front diffuser wall modified to change the characteristics of the turbocharger, FIG. 17 is an explanatory diagram of compressor performance for a conventional turbocharger, and FIG. The figure is a compressor performance explanatory diagram for the turbocharger shown in Figs. 1 to 16, Fig. 19 is a vertical cross-sectional view of a modified version of the turbocharger shown in Figs. 1 to 16, and Fig. 20 is a Fig. 16 is a vertical sectional view showing the turbocharger installed between the V-type machines, and Fig. 21 is a vertical sectional view showing the turbocharger installed between the V-type machines.
0 is a top plan view, FIG. 22 is a vertical sectional view showing the turbocharger shown in FIGS. 1 to 16 attached to an in-line engine, FIG. 26 is a top plan view of FIG. 22, and FIG. FIG. 25 is a vertical sectional view of the upper half of a turbojet equipped with a bearing support structure, FIG. 25 is a vertical sectional view of a turbofan equipped with the bearing support structure of the present invention, and FIG. 26 is a modification of the turbojet shown in FIG. 24. Vertical sectional view of the upper half of the example, No. 27
The figure is a sectional view taken along line 27-27+ll1l in Fig. 26, Fig. 28 is a vertical sectional view showing the upper half of the turbofan equipped with the bearing support structure of the present invention, and Fig. 29 is shown in Figs. 1 to 16. 30 is a vertical sectional view of a generator applied to a turbocharger, which can also be applied to the turbojet and turbofan shown in FIGS. 24 to 29. FIG. 206.210...First and second inner races, 204
・・・Fixed lace, 20 mouths ・・・Outer nozzle, 2
12... Outer race ring, 218... Spring means, 22
2224...Wick FIG, 9 PT2 / PTl 5, 4 FIG, 23 Procedural amendment band (discharge) Mr. Commissioner of the Patent Office 1, Indication of the case 1999 Patent Application No. 250809 3, Person making the amendment Related free time d morning sleep aspawn. There are no changes to the contents except for Norbat and Elle)” FIG, 29

Claims (3)

[Claims] (1) In a bearing support structure that supports rotation of a compressor and a turbine of a turbocharger, (a) a shaft extending from the compressor and the turbine and having first and second inner races formed therearound; (b) an outer housing having a fixed race corresponding to the inner race so as to receive a plurality of balls between it and the first inner race; (c) an outer housing relative to the fixed race of the outer housing; (d) a second outer race ring that is slidably slidable and positioned adjacent the second inner race to receive a plurality of balls therebetween; fixed lace,
the second inner race being disposed between the first inner race, the second outer race ring, and the second inner race to position the shaft relative to the outer housing; (e) spring means for biasing the outer race ring of the outer housing away from the fixed race of the outer housing; (e) formed on the shaft and disposed between the fixed race of the outer housing and the first inner race; (f) during rotation of said shaft, lubricant is applied to said peripheral slope and is carried by centrifugal force to said ball; lubricant means for supplying a lubricant to the peripheral slope through a wick having one end that contacts the lubricant and an opposite end that contacts the peripheral slope so that the lubricant is supplied to the peripheral slope. Support structure. (2) A lubricating means for supplying a mist of lubricant to the balls between the fixed race of the outer housing and the first inner race to lubricate them. The bearing support structure described in (1). (3) a second peripheral ramp formed on the shaft and angled toward the ball disposed between the second outer race ring and the second inner race; , one end in contact with lubricant, and the second periphery such that during rotation of the shaft, lubricant is applied to the second periphery ramp and carried by centrifugal force to the ball. A bearing support according to claim 1, further comprising lubricant means for supplying lubricant to the second peripheral slope via a wick having an opposite end in contact with the slope. Structure.