WO2010111133A2

WO2010111133A2 - Reduction of turbocharger core unbalance with balance washer

Info

Publication number: WO2010111133A2
Application number: PCT/US2010/027933
Authority: WO
Inventors: Denny King
Original assignee: Borgwarner Inc.
Priority date: 2009-03-26
Filing date: 2010-03-19
Publication date: 2010-09-30
Also published as: WO2010111133A3

Abstract

Turbochargers operate at extremely high speed, so balance of the rotating core is of the utmost importance to turbocharger life. A special balancing washer is added to the clamping region between the compressor nut and the nose of the compressor wheel to aid in keeping the wheel, nut, and stub-shaft on the turbocharger axis and to thereby reduce the degree of core unbalance.

Description

REDUCTION OF TURBOCHARGER CORE UNBALANCE WITH

BALANCE WASHER

FIELD OF THE INVENTION

This invention addresses the need for improved core balance throughput, and accomplishes this by designing a special balance washer.

BACKGROUND OF THE INVENTION

Turbochargers are a type of forced induction system. They deliver air, at greater density than would be possible in the normally aspirated configuration, to the engine intake, allowing more fuel to be combusted, thus boosting the engine's horsepower without significantly increasing engine weight. This can enable the use of a smaller turbocharged engine, replacing a normally aspirated engine of a larger physical size, thus reducing the mass and aerodynamic frontal area of the vehicle.

Turbochargers (Figs. 1 and 2) use the exhaust flow, which enters the turbine housing (2) from the engine exhaust manifold to drive a turbine wheel (51 ), which is located in the turbine housing. The turbine wheel is solidly affixed to the turbine end of a shaft, becoming the shaft and wheel assembly (50). A compressor wheel (20) is mounted the other end of the shaft, referred to as the "stub shaft" (56), and the wheel is held in position by the clamp load from a compressor nut (30). The primary function of the turbine wheel is providing rotational power to drive the compressor. Once the exhaust gas has passed through the turbine wheel and the turbine wheel has extracted energy from the exhaust gas, the spent exhaust gas exits the turbine housing and is ducted to the vehicle downpipe and usually to the after-treatment devices such as catalytic converters, particulate filters and NO_x traps.

The compressor stage is mainly comprised of a wheel (20) and it's housing (10). Filtered air is drawn axially into the inlet of the compressor cover by the rotation of the compressor wheel (20). The power generated by the turbine stage to the shaft and wheel drives the compressor wheel to produce a combination of static pressure with some residual kinetic energy and heat. In one aspect of compressor stage performance, the efficiency of the compressor stage is influenced by the clearances between the compressor wheel contour (28) and the matching contour (13) in the compressor cover. The closer the compressor wheel contour is to the compressor cover contour, the higher the efficiency of the stage. In a typical compressor stage with a 76mm compressor wheel, the tip clearance is in the regime of from 0.31 mm to 0.38mm. The closer the wheel is to the cover, the higher the chance of a compressor wheel rub, so there has to exist a compromise between improving efficiency and improving durability.

Viewed on an oscilloscope the wheel in a compressor stage does not rotate about the geometric axis of the turbocharger, but rather describes orbits roughly about the geometric center, as seen in Fig. 3. The geometric center (35) is the geometric axis of the turbocharger. The compressor end of the turbocharger, with data taken from a cylindrical nut, describes the orbit (81).

The dynamic excursions taken by the shaft are attributed to a number of factors including: the unbalance of the rotating assembly; the excitation of the pedestal (i.e., the engine and exhaust manifold); and the low speed excitation from the vehicle's interface with the ground.

As a dynamic assembly, the rotating assembly passes through several critical speeds. At the first critical speed, the critical mode is rigid body bending. In this mode the rotating assembly described a cylinder. At the second critical speed, the critical mode is again that of a rigid body, but in the conical mode about the outer ends of the bearing span. At the third critical speed the critical mode is that of shaft bending. The third critical speed occurs at from 50% to 70% of the speed range through which the turbocharger operates (namely, typically from about 30,000 up to 140,000 RPM). The first two critical speeds are much lower than that and are passed through very quickly during accelerations.

The first two modes are predominantly controlled by the bearing stiffness. The third mode, that of shaft bending, is predominately controlled by the stiffness of the shaft. The stiffness of the shaft is proportional to D₅ ⁴, where D_s is the diameter of the shaft.

The power losses due to the bearing system are predominantly controlled by D_s ³ so it can be seen that the control of the third critical mode is a compromise between power losses, thus efficiency and shaft bending. When there is an unbalance force acting on the rotating assembly at the compressor-end of the turbocharger, the stiffness of the shaft is a major factor in countering that force and also in allowing the turbocharger to continue to run after a compressor wheel rub against its cover.

After a loss of oil pressure or oil flow to any of the journal or thrust bearings, the predominant ultimate cause of turbocharger failure is contact between a wheel and cover. This contact can be as mild as a rub of the rotating wheel on the cover, or as serious as an impact of the wheel on the cover. To minimize the risk of this contact, the manufacturer takes many steps to build dynamic integrity into the rotating components. In a mid-sized, commercial Diesel turbo, for example with a 76mm compressor wheel, the shaft and wheel (50), seen in Fig. 2, which is recognized as the welded assembly of the turbine wheel (51 ) to the shaft, is balanced in two planes: The plane perpendicular to the shaft at the nose (89) and the plane perpendicular to the shaft at the backface (88). Since the shaft and wheel is finished as a very accurately machined, single component with shaft diameters ground to tolerances in the 2.5 micron regime, its inherent balance is quite good. In addition to these tightly held diametral tolerances, the diameters which support the journal bearings (70) on the large diameter end (52) of the shaft and the stub shaft (56), upon which the compressor wheel and smali parts are both axiaily and radially located, are held to a complex cylindricity tolerance measured in the regime of a micron.

The shaft and wheel component, for the turbocharger size above, is balanced within a range of 0.1 to 0.5 gm/mm.

The next components for discussion in the rotating assembly are the thrust washer and flinger. Both components are ground steel and of relatively small diameter when compared to a wheel. The thrust collar has a mass of around 10.5 gm; the flinger has a mass of around 13.3 gm. Because they are totally circular and have a high degree of finish, these components have very dose to perfect balance. The next component is the compressor wheel, which has a mass of around 199 gm.

The compressor wheel is an extremely difficult part to machine and balance. While it is ultimately balanced to a range from 0.1 to 0.5 gm/mm in each plane, getting down to that limit is difficult. It is extremely critical to machine the bore (27) in the center of the wheel such that it is centered on the hub at both the nose end face (21 ) and the hub end (22). This means that the majority of the mass of the machined wheel is centered on the bore (27) of the compressor wheel. The act of centering the as-yet un-machined casting on the imaginary turbocharger centerline (35) also results in blades of equal length, which further contributes to the balance of the component. If the wheel is not chucked exactly on center with the hub profile, the machining of the blade contour surfaces (28) off center (of the hub) results in blades of different lengths. Blades of unequal length can cause not only balance and blade frequency problems, but also once-per-revoiution unwanted acoustic probiems. The function of the compressor nut is to apply sufficient clamp load to the compressor wheel such that it will not rotate under any dynamic conditions, from max speed from cold start, to hot shutdown at max speed. However, in view of the influence of the nut on balancing, the compressor nut should not be referred to as a nut in the normal sense of the term. While the nut is a relatively low mass item, at 6.3 gm in the turbo under discussion, its contribution to unbalance (as against balance) can be very large. A requirement of the nut is that the lower face (31 ), the face in contact with the face on the nose end face (21 ) of the compressor wheel, must be manufactured to a very tight perpendicularity tolerance to the bore of the thread in the compressor nut, in the range of 0.03 to 0.04mm, so that when the nut is threaded onto the shaft, and clamp load applied, the aforementioned lower face of the nut is applying a load close to normal to the face (21 ) on the nose of the compressor wheel. Failure to apply this load either normal to the face of the compressor wheel, or parallel to the shaft centerline (35), will cause bending of the shaft, with the result that the mass of the compressor wheel, nut, and stub shaft will be displaced from the turbocharger axis (35) causing a large unbalance in the rotating assembly. Since the nut is extremely difficult to assemble exactly on axis, the mass of the nut is a critical factor in the level of unbalance the bearing system can tolerate. For the same degree of unbalance in the core, the lower the mass of the nut, the higher the geometric run- out acceptable tolerance. Much effort goes into the design of the top of the compressor wheel (21), the nut (30), and the amount of thread (57) visible above the nut to keep the mass in this zone to a minimum. If the nut is not perpendicular to the top of the compressor wheel, and parallel to the stub shaft below the nut, then the threaded part of the stub shaft above the nut (i.e., with thread no longer engaged with the thread on the stub shaft) wiii also be off-center with the centerline of the stub shaft beiow the nut and ultimateiy off-center with the turbocharger axis, thus contributing to even greater core unbaiance.

At the point of manufacturing, all of these critically balanced items are assembled and the core balance, that is the balance of the rotating assembly, assembled to the bearing housing, supported by the journal bearings, is spun at high speed, with oil pressure supplied to support the rotating shaft on its designed oil film. This procedure checks the balance of the rotating "core". If the balance is within limits, then the core is satisfactory and is released for assembly into a complete turbocharger. If the balance is out of limit, then the core undergoes a procedure to bring the balance into limits before it is assembled into the housings to produce a turbocharger.

Accordingly, when the turbocharger leaves the factory, the rotating core is within a balance limit, and the turbocharger could be expected to live for several engine rebuild periods.

However, in the period the turbocharger is operating on the engine, the balance of the rotating core can be degraded in many ways, some of which are listed here: the turbine wheel is subjected to damage from particles, sometimes quite large, from the combustion chamber and, in case of EGR, the exhaust manifold, which causes damage ranging from bending to breaking off of parts of the blades, which then causes a deviation from the factory balance condition; the compressor wheel also can be subjected to damage inflicted by "foreign objects" which are ingested into the system. Loss of oil pressure for a period can cause loss of support of the rotating assembly, which can result in a wheel rub on either or both wheels, which, at minimum, can cause the removal of some blade material (by rubbing on the housing), which then alters the mass of several adjacent blades, or in a heavier rub, can bend the blades. Both of these events may cause a change in the balance of the rotating assembly.

If the rotating assembly does develop an unbalance condition less than those discussed above, the resultant of the core unbalance can be the generation of noise at a once per revolution frequency. With a turbocharger rotating at 150,000 RPM to 300,000 RPM, an unbalance related acoustical event will be in the frequency range of 2,500 to 5,000 Hertz. This makes the frequency somewhere around the highest producible by a flute (2093Hz) and the highest producible by a piano (4186 Hz) so the customers do complain about the noise.

A measure of the efficacy of a turbocharger bearing system is the ability of the bearing system to control and support the rotating assembly under all conditions. Turbocharger bearing systems come in many designs, from ball bearings for very large and some high performance turbochargers, to different configurations of fixed sleeve bearings, floating oil film bearings and air bearings. They ail have one thing in common, and that is the need for fine balance control of the rotating assembly. The level of balance for the individual components is generated, to some extent, by the level of balance acceptable by the bearing system in the rotating assembly. An automotive type, oil pressure fed, well designed bearing system will present to a manufacturer a maximum unbalance which the bearing system can control and which will provide sufficient damping that it remains in control of the shaft excursions under ail conditions. This means that any balance condition under the maximum unbalance condition acceptable for that bearing system on a specific engine is acceptable from an engineering point of view. The cost to achieve this level of core unbalance increases as the level of acceptable unbalance decreases. In the experience of the inventor, some turbocharger cores pass through the core balance "gate" with no additional attention. Some cores need attention, which can be as little as undoing the compressor nut, rotating some components, re-applying the clamp load, and then re-testing, to replacing components in the rotating core.

The goal of a turbocharger manufacturer is to offer product at the lowest cost, with the highest possible reliability and durability. Balance is a key factor in the durability and reliability facets. So it can be seen that there is a general need to present cores to the core test device which fall well inside the unbalance lower limit in an effort to both decrease assembly costs and increase turbocharger life.

SUMMARY OF THE INVENTION

The above objectives were accomplished, and the present invention achieved, by the development of a special washer located between the interface of the compressor nut and the top surface top surface on the nose of the compressor wheel of the compressor wheel of a turbocharger, which reduces the standard deviation in incoming core balance turbocharger cores by over 43% while reducing the mean value by 58%.

In terms of production viability, which takes into account the minimum acceptable values and all turbo part numbers tested, the uncorrected or first-pass throughput of acceptable cores without the special washer was 22% while the throughput of acceptable cores using the washer was 42%, an increase of 90%.

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention is illustrated by way of example and not limitation in the accompanying drawings in which like reference numbers indicate similar parts, and in which:

Fig. 1 depicts a section of a turbocharger assembly;

Fig. 2 depicts the rotating components in a turbocharger;

Fig. 3 depicts the orbits made in testing;

Fig. 4 depicts the orbits of individual components;

Fig. 5 depicts a machined compressor wheel section; Fig. 6 depicts the compressor wheel mounted on a shaft;

Fig. 7 depicts the assembly of Fig. 6 subjected to runout of the nut;

Fig. 8 depicts the first embodiment of the invention;

Fig. 9 depicts some assembly statistics;

Fig. 10 depicts a histogram of the assembly unbalance; Fig. 11 depicts the second embodiment of the invention; and

Fig. 12 depicts a magnified view of the second embodiment of the invention.

DETAILED DESCRIPTION OF THE INVENTION

Turbocharger assemblies are core balanced to ensure required life and to control rotational vibration induced noise. The inventor realized that a high percentage of newly assembled turbocharger cores were not passing the core balance checking station, which meant that the turbochargers had to be re- processed, some, several times, to achieve a "pass" under the core balance limit. This re-processing resulted in both high processing and capital costs.

Contrary to the conventionally accepted design direction of reducing rotating mass and inertia, the inventor added mass and some inertia by adding a specially ground and hardened washer located between the compressor nut (30) and the top surface on the nose (21 ) of the compressor wheel. This washer prevents the nut from rocking and tracking on the nose of the compressor wheel.

As shown in Fig. 7, as clamp load is applied to the compressor wheel by rotating the nut to travel down the helix angle of the thread, several events can happen. The act of rotating the nut against the face (21 ), on the nose of the compressor wheel, can cause the nut to dig into the face and track off center, particularly when the nut is steel and the compressor wheel is aluminum. This tracking causes the mass center of the nut to move off the turbocharger axis, which results in an unbalance (N), equal to the mass of the nut times the displacement (R_n), perpendicular to the turbocharger axis.

This displacement also causes a bending of the stub shaft, which results in yet another unbalance force (S) in the same direction, which is equal to the mass of the stub-shaft (57) deviated from the turbocharger axis (35) times the displacement (R_s). The bending of the stub-shaft can also cause a displacement of the compressor wheel center-of-gravity, which is indicated in Fig. 7 as an unbalance force of "C". Resisting these bending events is the interaction of the surface of the stub-shaft (67) which is a sliding fit to the surface (26) of the hole (27) in the compressor wheel (20), aided by the compression of the clamp load applied by the interaction of the internal threads (32) in the compressor nut (30) against the threaded end (57) of the stub-shaft (56).

By adding the inventive balance washer (36) as shown in Fig. 8 to locate between the compressor nut lower face (31) and the compressor wheel nose surface (21 ), the nut is held true to the turbocharger axis (35) and the unbalance forces of the nut (N), the stub shaft (S) and the compressor wheel (C) are minimized or non-existent. As a result, the major unbalance force on the compressor end is confined to the imbalance of the compressor wheel component.

Fig. 9 depicts test results of core balance as measured without the special balance washer as well as with it. The sample group (106) depicts the core unbalance values for a set of cores which are presented to the core balancer directly from production. The value of three standard deviations is represented on the upper limit by the broken line (100), and the lower limit by the broken line (103). The broken horizontal line (109) represents the maximum acceptable core unbalance. That is to say that any cores with values beneath the broken line are acceptable; any above are un-acceptable in terms of core balance. In this sample set, none of the cores presented to the balancer are acceptable. The center data set (107) is the same set of cores, but this time fitted with the balance washer. Again the horizontal broken lines (101 and 104) are the upper and lower three standard deviation boundaries. In this case, 60% of the cores were acceptable. In the sample set to the right of the chart, the same set of cores was presented with the special balance washer removed. The range, displayed by the 3 standard deviation limits (102 and 105), has reverted to values close to those of the first case without the special balance washer. It should be noted, although, that some of the cores were just under the acceptable limit line. The histogram in Fig.10 shows the statistical curve (91 ) for a group of cores using the standard existing configuration and the statistical curve (90) for a group of cores using the special balance washer. The "X"-axis depicts the range of core unbalance, and the Υ"-axis depicts the density of results. The histogram shows that statistically the addition of the special balance washer not only moves the statistical curve (90) for the mean unbalance force closer to the left, which is a lower unbalance value, but it also groups the cores more tightly around the mean. Both of these features are positive in terms of throughput and reduced average unbalance values.

In the second embodiment of the invention, the special balance washer is designed such that the special ground and hardened washer (37), in Figs 11 and 12, is located between the compressor nut (30) and the top surface on the nose (21 ) of the compressor wheel. In addition to preventing the nut from rocking and tracking (i.e., rocking off the perpendicuiar (to the CL) face and then tracking on top of the CW as the edges dig in) on the nose of the compressor wheel, the flanged component centers the compressor wheel on the washer, and the inner diameter of the flanged balance washer radially centers the assembly on the shaft at the nut end of the stub shaft.

In the second embodiment of the invention, the compressor wheel, at the backface (22) or lower end, has a short section of the bore surface (26) which is a sliding fit on the stub shaft outer surface (61). This zone, with the inner diameter (45) of the flanged balance washer (37), creates the radial location of the compressor wheei, with respect to the turbocharger axis (35), without the need for the bore of the compressor wheel to maintain cylindricity for its entire length. The undercut (29), which becomes clearance from the stub shaft outer diameter surface, (61) can be the location for the outer diameter (46) of the flanged section diameter, or the flanged section of the special balance washer can fit in a counter-bore (33) in the compressor wheel nose.

The special balance washer has an added benefit, outside the reduction in core unbalance, and that is providing a consumable mass for balance correction to the balance correction effort. Normally the balance correction (25) is executed by machining aluminum from the side of the nose of the compressor wheel as seen in Fig. 6. With a steel special balance washer, the mass removal for balance fine tuning can come from the removal of steel (which has higher density than aluminum). With a steel washer, less volume of material has to be removed for the same mass correction with still another benefit in that, when the compressor wheel is used for a second or third life there is still material available for balance correction.

Now that the invention has been described, I claim:

Claims

1. A rotating assembly, comprising a shaft (52) with a turbine end and including a reduced diameter stub shaft (56) part with a threaded end, a turbine wheel (51 ) rigidly connected to the turbine end of the shaft to form a shaft-and-wheel assembly (50), a compressor wheel (20) held in position on the stub shaft (56) by the clamp load from a compressor nut (30) threaded onto said threaded end of the shaft, and a balance washer (36, 37) provided between the nut and the compressor wheel.

2. The rotating assembly as in claim 1 , wherein the balance washer (36, 37) is made of a material having a higher hardness than the material of the compressor wheel.

3. The rotating assembly as in claim 1 , wherein the stub shaft (56) has an outer diameter, the compressor wheel (20) has a nose end face (21 ), a hub end face (22) and a bore (26) with an internal diameter at the nose end face greater than the outer diameter of the stub shaft at the compressor wheel nose end face, and wherein balance washer (37) includes a part extending radially and flush with the compressor wheel nose end face, and an axial cylindrical flange part (45) extending into the bore between the compressor wheel internal diameter and the shaft outer diameter, wherein the inner diameter (45) of the flanged balance washer (37), creates the radial location of the compressor wheel, with respect to the turbocharger axis (35).

4. The rotating assembly as in claim 3, wherein the compressor wheel, at the hub end (22), has a bore surface (26) which is a sliding fit free of play on the shaft outer surface (61 ).

5. The rotating assembly as in claim 1 , wherein the shaft has an outer diameter, the compressor wheel (20) has a nose end face (21 ), a hub end (22) and a counter-bore (33) at the nose end face, and wherein balance washer (37) includes a part extending radially and flush with the compressor wheel nose end face, and an axial cylindrical flange part (45) extending into the counter-bore (33) between the compressor wheel and shaft.

6. The rotating assembly as in claim 1 , wherein the compressor wheel (20) is made of aluminum or an aluminum alloy.

7. The rotating assembly as in claim 1 , wherein the compressor wheel (20) is made of titanium or a titanium alloy.

8. The rotating assembly as in claim 1 , wherein the compressor wheel (20) is made of a ferrous metal or ferrous alloy.

9. The rotating assembly as in claim 1 , wherein the balance washer (36) is flat.

10. A method for balancing a rotating assembly comprising a shaft (52) with a turbine end and including a reduced diameter stub shaft (56) part with a threaded end, a turbine wheel (51) rigidly connected to the turbine end of the shaft to form a shaft-and-wheel assembly (50), and a compressor wheel (20) held in position on the stub shaft (56) by the clamp load from a compressor nut (30) threaded onto said threaded end of the shaft, the method comprising:

(a) introducing a balance washer (36, 37) between the nut and the compressor wheel, (b) spin testing the rotating assembly for balance within a predefined limit,

(c) in the case that the rotating assembly does not pass the balance test, machining away a part of at least the balance washer, and

(d) repeating steps (b) and (c) until said rotating assembly passes said balance test.