Effect of applied moment on B10 life? 2

[jstewart]

I am working on a small bearing application (20mm shaft) where the load is primarily radial, with a very small amount of overturning moment. There's not much axial space to work with, and the main candidates in my design are currently a deep-groove ball or needle-roller bearing with a thrust washer set.

The ISO and ABMA standard calculations for B10 life include factors for calculating an "effective" radial load for use in the life equation, but say nothing about applied moments.

Does anyone know of a good way to factor those loads into the calculation?

 

[dimjim]

Resolve the moment loading into
thrust and radial components.

 

[jstewart]

Sorry, perhaps I wasn't too clear. If there were two bearings, I could resolve the moment into two radial loads. But in my application there's just a single bearing under load. Thus, that bearing must support the small-but-nonzero moment load.

To my knowledge, moment loads on single bearings cannot be resolved into purely axial and radial components. I'd be happy to be proven wrong, though.

I know that deep groove ball bearings can support some small moment loading; I want to know how much it will affect the B10 life.

 

[JeanMicheling]

I'd say that you should not have only one bearing to withstand moment load. The tilting effect will change your contact angle and considerably reduce the BSL. You should contact bearing manufacturers and see what they have to say. I'd be curious to know what they think about that.

 

[jstewart]

Ideally 2 ball bearings would be used to take out the moment, but as I mentioned, the axial space is limited, and the moment is quite small. I've seen competitive designs use a similar arrangement with good results.

I'm still waiting to hear back from the manufacturer. The problem was too complex for the local applications "engineer"/salesman, so it's been referred to the guys who do the calculations. I just thought I'd ask here and see if anyone had run into a similar problem.

Also, I'm starting to look into codes (software) for bearing analysis (looking at ball loading, hertz stress, internal clearance, effect of bearing fit-up), and was wondering if anyone had any recommendations for a good basic package? I've looked at COBRA-AHS/EHL so far. Can you even buy A.B. Jones anymore? I don't need any rotordynamics stuff; quasi-static is fine for what I'm doing.

Any ideas?

 

[EnglishMuffin]

Try this :

http://www.advanced-bearing.com/

TK solver based - not cheap - used to cost about 5 grand.

 

[JeanMicheling]

FAG seems to have a really great software as well. I never tried it but it seems to be interesting. Go on the website for more details.

 

[jstewart]

The TK-Solver based software looks pretty comparable to the COBRA-AHS "baseline" package. (price and features) Pretty good, but maybe more than I need for my application. COBRA-EHL might be a better fit for me.

It's interesting that FAG/INA has decided to provide their Bearinx software as an online service. It looks like a very capable package, but unfortunately my company doesn't buy direct from INA (you have to be a direct account to access Bearinx.)

I found some moment calculation information in Harris' "Rolling Bearing Analysis." Suffice it to say that the method is not simple. Maybe software is the right way to go with this one.

 

[EnglishMuffin]

I believe you can still get Bert Jones' (the original) - I think somebody told me that his son sells it, but I'm not sure. When ABODE first came out, I think it was then the best available, other than bearing manufacturer's proprietory stuff. But the top of the line COBRA is probably the best now - partly developed by Poplawski on government grant money. It has flexible shaft - ABODE doesn't. There are also free copies of the venerable SHABERTH floating around if you can find them - old mainframe software - very unstable and hard to use.

 

[ischgl99]

jstewart, INA does keep a tight control over who has access to the Bearinx program, the applications engineers at INA/FAG should be able to do the analysis for you fairly quick. It looks like you are going to have a problem with misalignment in this application if you use a deep groove ball bearing unless you have some form of constraint to resist the moment. Would you have space for a double row angular contact bearing? They can handle the moment loads.

 

[jstewart]

I'd consider submitting this problem to INA/FAG, but my company doesn't buy ball bearings from them, just roller clutches. That's not likely to change anytime soon, and even if I could, I'm not comfortable taking advantage of their "free" service for a product I don't buy.

I am perfectly willing to pay for analysis and related tools, but I don't see that INA/FAG has made any provision for that.

My current supplier will eventually respond (they're a quality manufacturer also), but if I try to iterate the design through them, it could take weeks.

mbensema, there might be room for a double-row angular contact bearing, but I think this is probably overkill. I've also considered a 4-point contact bearing, which should be able to handle more moment due to the increased contact angle.

But again, the moment loading is very small compared to the radial load, and should be well within the performance envelope of even a basic radial ball bearing. The problem is still how do you calculate B10 life on any of these bearings with this loading?

I suppose I could code up the equations from Harris/Jones, but it's probably better to just buy an off-the-shelf program, since they have more capability and have already been debugged and proven.

Arg.

 

[dimjim]

I think your question is basic.
You might want to check out Tedric Harris's
books on Roller Bearing Design Anaylysis.

Even if you have a program, you will need
the basic bearing parameters:
Bearing Diameter
Ball Diameter
Contact Angle
Conformity Radius
Amount of Clearance or Preload

Harris does have equations to determine
the effects of moment and combined loads.
Almost all of the bearing formulas assume
use the most heavily loaded ball or roller
to base their calculations.
If you can resolve the effect of the moment
on the top ball tdc, I think you can use
the basic equation that the b10 life will
decrease by the ratio of the third power
of the loads.

 

[dimjim]

As an example:
Assume that the ball load increases by 10 percent
1.1 cubed would be 1.331. The reciprocal would
be 75 percent of the b10 life.

 

[volpegrigia]

There is a code that was sold by the Franklin Institute in Philadelphia. It was called Genrol. It was a fortran code that was sold with the source. It was/is quite good. I worked for INA for 26 years (I am no longer there) and we had Genrol which has now been supplanted by the in-house developed BearinX software (which is quite good). I can run your analysis if you are interested. I have a different computer code (similar to BearinX but not commercially available and without the graphic output and input and I can get you an answer quickly. Just think in this term. If the radial load is quite high with respect to the moment, then the impact is small. If the opposite is true, then the impact of the moment may be significant.
If you send me your conact information I will send you a couple of papers on the subject and you can write your own mini-program. Since you have ted's book, you may want to check the section with the moment load. It appears more complicated that it actually is.

 

[Tmoose]

If cleanliness is good and ElastoHydroDynamic lubrication is achieved then B10 life can be a vastly pessimistic estimating tool. Or, if turned around the other way, an empty promise when used to predict important bearing life improvements as a result of a few trim balance weights.

 

[volpegrigia]

I have calculated a bearing 6204 (20 mm bore radial bearing with 0.008mm clearance) and the equivalent load changes as follows:
Fr=radial load; M=moment; P&Po=equivalent loads (dynamic & static).
Fr=100N; M=1Nm; P=234N Po=222N
Fr=200; M=1; P=320; Po=318
Fr=300; M=1; P=408; Po=414
Fr=500; M=1; P=595; Po=605
Fr=1000; M=1; P=1080; Po=1080

If you change the clearance the equivalent load is affected as well. By doubling the clearance to 0.016mm, the equivalent load decreases as follows:

Fr=100; M=1; P=199; Po=192
Fr=200; M=1; P=296; Po=294
Fr=300; M=1; P=390; Po=390
Fr=500; M=1; P=580; Po=585
Fr=1000' M=1; P=1070; Po=1070

You can draw your own conclusions.