Here's part of a Machine Design article I found on Dynaroll's website:
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Regardless of component design, bearings need preload to remove axial
play and boost axial and radial stiffness. Applying preload to inner
rings of bearing pairs maximizes stiffness. Conversely, loading outer
rings lowers stiffness. There are three basic preload methods: springs,
solid clamping, and deadweight. Springs press together inner rings or
force apart outer races. Springs ease assembly though system stiffness
suffers because spring rate rather than raceway-ball elasticity controls
stiffness. Springs also minimize relative thermal expansions between
mating parts, important for assemblies of different materials exposed to
temperature extremes. High-speed applications typically use spring
preload as well. For Belleville or wavy-washer springs, tolerance stack
up of mating parts can nearly equal maximum allowable compression of the
springs themselves. Use a low-rate spring to minimize preload variations.
Solid clamping removes axial play by machining components to precise
dimensions. The method produces stiff assemblies and cuts component
count. However, normal tolerances in bearing axial play make it
difficult to do in practice. Components and bearings must be tolerance
matched with high precision or bearings may suffer raceway brinelling
(overloading) or have insufficient preload. Precision-matched, duplex-pair
bearings solve the tolerance problem but cost significantly more than
standard bearings.
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I'm a bit confused by that because if you pull the inner races together,
balance of forces mean the outer races will be pushed out by an equal
force. I'm not really worried about bearing life though because this
isn't a continuous rotary application. It will probably be actuated on
the order of a 100-200 times max.
I agree with your 3rd point that to pull the inner races together, the
outer sleeve needs to be slightly longer than the inner one. I screwed
that up so thanks for catching that. I don't think there will be any
extra load on the ball because 1) the shaft/stem is not pressed against
the slot in the ball, there is some clearance there, and 2) the internal
pressure will tend to push the shaft out. I believe the ball will stay
put since the body seals are contoured to fit the ball.
I'm not sure I understand your last paragraph. The only reason I used a
flanged bearing at the top is to anchor it in place between the valve
body and retainer plate so everything is relative to that top bearing.
Otherwise, when not pressure loaded, the stem could be forced down
slightly and crush the seal.
Thanks!
-Bob
On 05/22/2016 08:11 PM, Norman Yarvin wrote:
I'll third the Belleville washer; this thing will be seeing a fair bit
of thermal expansion and contraction, so a spring washer to keep the
preload from varying too much is a good idea.
Not that I've really figured out what the idea is with the preload.
Looking at your diagram:
http://www.watzlavick.com/robert/rocket/rocket1/drawings/ball_valve_3_assy-annotated.pdf
preload could come from:
1. Compressing the inner races vertically, via the preload
nut, whereupon they bulge out horizontally (per Poisson's
ratio).
2. Press fitting the bearings onto the shaft and into the
hole, which you've been doing from the start but obviously
wasn't enough (though probably would have been with a really
hard press fit),
3. Forcing a bearing's races in opposite directions (on the
diagram, one up and the other down), which you may be doing
some of, via the flange on the top bearing being opposed to
the valve ball, and via the two bearing spacer sleeves being
slightly different lengths. But to get preload on the bottom
bearing via that last, you'd want the outer sleeve longer,
not the inner one... though that is probably how things end up
once the preload nut is tightened and the whole stack gets
compressed.
You'll also get some of #3 from thermal expansion. As a rough figure,
with the housing being aluminum and the stem stainless, the difference
in thermal expansion coefficients seems to be about 5 parts per
million per degree C; with 200 degrees C temperature change, that's 1
part per thousand.
But as regards preload coming from the flange on the top bearing being
opposed to the valve ball (the flange pulling the outer race up, and
the ball, via the preload nut, pulling the inner race down), that is
the most vulnerable to thermal expansion of these sorts of preload,
and also the least desirable, since it adds to the friction of the
valve ball. Doing a quick search on those part numbers, the bearings
you've specified are "deep groove"; there are also "shallow groove"
bearings which allow more of a vertical float. (Not that deep groove
bearings will necessarily be a problem here -- in fact, I doubt they
will, since the dimensions are relatively small -- but if they do
cause problems, there are alternatives.)