Static unbalance-type balance shafts with axis alignment...

Details Static unbalance-type balance shafts with axis alignment... Static unbalance-type balance shafts with axis alignment...

2002-01-14

2004-01-27

Browne, Lynne H. (Department: 3679)

Rotary shafts, gudgeons, housings, and flexible couplings for ro

Miscellaneous

C123S192200, C074S603000

Reexamination Certificate

active

06682437

ABSTRACT:

TECHNICAL FIELD
The present invention relates to balance mechanisms for rotating machinery, particularly balance shafts for multi-cylinder internal combustion engines which exhibit vertical shaking forces.
BACKGROUND ART
Balance shafts are commonly used to reduce or cancel shaking forces and/or vibrations which result from residual imbalances inherent in the design architecture of machinery with rotating and/or reciprocating parts, or mechanisms, such as motors. These balance shafts are sometimes called “counterbalance” shafts.
Balance shafts are particularly valuable when operator or passenger comfort and freedom from noise and vibration related fatigue or distraction are desired, as in the case of motor vehicles such as automobiles, motorcycles, and the like. It is also advantageous to minimize vibration from the standpoint of equipment reliability. Where vibrations are reduced, the size, mass, and/or complexity of the mounting structures can often be reliably reduced, thus potentially reducing costs.
With multi-cylinder motor vehicle engines, the inline four-cylinder engine configuration is favored in much of automotive and industrial use today due to its inherent packaging space, manufacturing cost, and fuel consumption efficiencies. These engines benefit from Lanchester-type balance shafts, which can cancel nearly all of the inherent twice-per-revolution shaking forces produced by this otherwise mass-balanced architecture.
Lanchester-type balance shafts for these inline four-cylinder engines are paired to rotate in opposite directions at twice engine speed. The two balance shafts are timed to cancel each other's lateral forces while opposing the vertical “secondary shaking forces” that result from connecting rod tilt, causing piston motion to depart from sinusoidal or “simple harmonic” motion in the midstroke region centered about 90 degrees before and after “top dead center.” Each shaft produces a single, or “static” rotating unbalanced or centrifugal force, which taken together with its mating shaft's rotating unbalanced force, produces a resultant vertical shaking force which most effectively is located centrally among the bank of cylinders so as to be coincident with the engine's resultant shaking force. Static unbalance-type shafts of this general type are shown, for example, in U.S. Pat. No. 5,857,388.
Helical gears are often employed as the means of maintaining orientation “timing” between Lanchester-type balance shafts because of their potential to represent the best value choice for engineering priorities such as durability, wear resistance, noise emissions, cost, packageability, mass, power consumption and the like. This potential, especially in the often critical case of noise emissions, is highly dependent upon actualization of the gearset's theoretically possible total contact ratio (TCR). This, in turn, is much more highly dependent upon freedom from axial misalignment of gear tooth geometries, i.e., gear tilting or twisting with respect to its mate, throughout the operating speed range, than are alternative timing drive means such as chains or toothed belts.
Additionally, it is often advantageous to use a driven balance shaft to drive another component, such as an oil pump, as taught in U.S. Pat. No. 5,918,573 entitled “Energy Efficient Fluid Pump,” to provide a synergistic noise control benefit. In this case, also, the stability of the driving means' axis alignment throughout all operating speeds can play a critical role with regards to many of the same engineering priorities mentioned earlier. Given the elastic compliance inherent to structural materials, operating shape changes will accompany the load changes that occur as a balance shaft is operated at various speeds. Adding material to increase the section modulus of structural members, such as connector portions of balance shafts, carries a mass (and often cost) penalty, and can carry a fuel consumption penalty and/or packaging space penalty as well, and yet can still not succeed in providing for engineering target value axis alignment control.
Accordingly, there exists an advantage in the ability to achieve targeted axis alignment control for balance shaft drive means such as gears and extensions throughout the operating speed range without incurring additional mass, cost, fuel consumption or packaging space penalty.
With these motivations in mind, an example of a typical balance shaft for an inline four-cylinder engine is shown in a purely static condition in FIG.
3
. This prior art balance shaft
10
includes counterweights
12
,
14
located adjacent to either side of (or “straddling”) the principal balance shaft journal
16
so as to apply their composite centrifugal loading to the journal bearing
16
, thus transmitting the centrifugal loads to the engine's structure. This arrangement, with the diameter of the bearing journal being smaller than the effective diameter of the counterweight (as defined by the locus of the counterweight's largest effective radius as the shaft rotates), represents the combination of power-consuming bearing friction and space-consuming unbalance mass best able to maximize fuel consumption, mass, and packaging space efficiencies in most cases. The bending stiffness of the ideally sized journal is typically lower, however, than that of a larger diameter, suboptimal (in terms of friction and heat generation) journal configurations. Journal bending stiffness typically plays a significant role in operating shape stability versus operating speed, yet can be utilized to advantage without penalty using the inventive strategy disclosed herein.
Referring to
FIG. 4
, which illustrates the balance shaft
10
with high speed operating shape that is greatly exaggerated for clarity, the counterweights
12
,
14
of this prior art configuration are “balanced,” in terms of their bending moments, the product of their respective centrifugal force magnitudes and locations, about the midpoint of the length of the principal journal
16
such that their composite resultant force is located at the midpoint
18
of the journal
16
when the shaft
10
is spun about its axis of rotation C
L
,
30
. By locating the composite resultant force at the journal midpoint
18
in this fashion, the balance shaft's output loads are typically located at the midpoint of the engine's cylinder array, which is typically the midpoint of its central bulkhead's axial length. This is accomplished by distributing the counterweight masses such that their “moments” of unbalance sum to zero at the midpoint
18
of the journal
16
, which is the targeted Effective Plane of Static Unbalance (“EPSUB”) location. This is discussed in more detail in U.S. Pat. No. 6,237,442, entitled “High Value Static Unbalance-Type Balance Shafts.”
In other words, the product of unbalance magnitude and its effective distance from the EPSUB is the same for each counterweight
12
,
14
, with the “effective distance” being measured from the center of gravity (CG) of each unbalance mass, respectively. These opposingly “balanced” bending moments act to bend the shaft structure principally in the region of the principal journal
16
, tending to deflect both ends of the shaft
10
away from its axis of rotation C
L
,
30
toward the CGs of the unbalance masses at elevated rotational speeds. The end result of this bending acts to result in a tilt, or “wobble” (lateral runout) of the drive gears
20
,
22
whose quiet operation depends heavily on axis alignment with respect to each other, i.e., the avoidance of such tilt. This unwanted gear tilt is generally represented by reference number
38
.
The mechanism which results in this tilt is an axial shift in the effective plane of support to the principal journal
16
provided by the sleeve bearing's oil film, as generally indicated by reference number
24
. This axial shift is a result of any tilting, or misalignment (e.g., due to high speed bending), of the principal journal
16
with respect to its bearing sleeve. Since the balance shaft&

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