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Main bearing potential to damp torsional vibration

Main bearing potential to damp torsional vibration

Main bearing potential to damp torsional vibration

For decades Many engines have sported torsional dampers on their cranks' noses.

Older Chevy sixes had main bearing journals all vary in diameter. The journal near the flywheel is largest, and going forward each journal is about 0.03" smaller than the previous one.

In a recent discussion an acquaintance proposed that this diameter variation results in some level of torsional damping for free that would not exist if all the main journals were the same size. I think part of his argument is a claim the smaller shaft diameter results in greater shaft deflection and lower surface speed, with the result the small journals run more eccentrically in the bearing thereby generating circumferential pressure differentials of greater magnitude within the fluid film, which he felt would result in "damping."

I know models of rotor dynamic analysis of turbines credit hydrodynamic bearing allow including a variety of damping factors and coeffients, but as best I recall they are typically radial damping, not torsional.

My questions - Is viscous torsional damping from crankshafts winding/unwinding < 1 degree ignored in engine design, since there is going to be a big old damper on the snout anyway?
And in any case, does that exist to a significant degree?

RE: Main bearing potential to damp torsional vibration

I'd argue the opposite effect: that the shear drag from larger journals would exhibit a greater damping effect (higher surface speed, thus more first-order drag for a given increase in rotational speed), and that the larger inertia of the larger crank bearings, would tend to make bigger journals more well-damped.

RE: Main bearing potential to damp torsional vibration

Take it a step further, your acquaintance is arguing that the bearing deflection creates locally higher drag on one side of the bearing, but it would also decrease on the opposite side, for a net change of...where's my fluids book...I think it's zero, but need to confirm.

RE: Main bearing potential to damp torsional vibration

I have tuned several production dampers,and at no point did we bother accounting for that effect specifically. It'd be easy enough to test for. If you think about it there is already a good reason for increasing the journal size towards the flywheel.


Greg Locock

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RE: Main bearing potential to damp torsional vibration

One good reason for stepping the bearing sizes is that the bearing bores can be machined all at once with a long bar, which can then be removed axially, even while still rotating, without danger of scratching any of the finished surfaces.

Mike Halloran
Pembroke Pines, FL, USA

RE: Main bearing potential to damp torsional vibration

Yeah, my guess was the variable journal diameters was a manufacturing driven.

Chevy used it at least as early as 193X on the first [three main bearing] Chebbie 6. They pumped up the main diameters pretty rapidly at first, but kept the stepped journals when they increased the number of main bearings right thru 1963.

RE: Main bearing potential to damp torsional vibration

I think it is likely that this was done to distribute load from the crank bearing to the block over a larger area on the loaded end of the crank. It also allows for more easily-maintained/manufactured clearances on those journals. It is possible that this dampens certain excitations, but reducing the bearing diameter in relation with distance from the loaded end of the crankshaft may have been to reduce drag where it didn't pay to have bigger load-bearing surfaces. I'd like to see some data on end-play on these cranks while under load.

RE: Main bearing potential to damp torsional vibration

Hi Panther140,

Why do you feel the flywheel end of the crank is "the loaded end?"


Dan T

RE: Main bearing potential to damp torsional vibration

The flywheel has a clutch and a transmission bolted onto it. The cylinders closest to this will transmit their torque with less crankshaft length between the flywheel and the cylinder during combustion.

This means that there will be less torsional flex from the crankshaft to dampen the impulses from combustions that are closest to the flywheel.

Picture a lever being used to lift a weight. The main bearings are the fulcrum in this lever. The load that the clutch couples to the flywheel represents the dead weight that we need to lift. The combustion represents the force we apply to this lever in order to lift the dead weight.

The combustions near the flywheel deliver a more concentrated impulse to the bearing, because it is not being dampened by torsional flex in the crankshaft. Distributing that impulse over a larger surface in the engine's case allows you to control where the stress is distributed, without making the case so rigid that stress is significantly more concentrated to the crankshaft.

That was the most simplified version of this concept. There are more sources of load, such as the rest of the crankshaft that the combustions still need to spin. Then there are pumping losses to deal with, which leads into the subject of engine braking. I haven't even touched on end-play or bending forces yet.

The combustions that happen on the side farthest from the flywheel are also dampened by increased torsional flex from the length of crankshaft which couples those combustions to the load. Reducing those bearing sizes serves to decrease drag and rotational mass, which also helps the PTO-side main bearings survive.

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