For lathe spindle bearings, i know one good option is to use two angular contact bearings on the far left, and a plain bronze taper bearing on the right. I was thinking about spreading the angular contact bearings apart to both ends of the headstock (while keeping a back-to-back arrangement and preload) and eliminating the bronze bearing. Is there any obvious reason why this would be problematic or inaccurate?

Bearings: ABEC 1 with 40 degree contact angle


For the spindle, i'm thinking about making it out of mild 1080 steel. If a plain bronze taper bearing is used, is 1080 steel compatible in terms of wear with the bronze bearing (assuming the spindle is polished to an optical finish with 0.3um and 0.05um polishing powder). Does anyone know what spindles are typically made out of (hardened and ground steel?)?

Heat buildup in the spindle lengthens it. The longer the distance between bearings, the greater the accumulative effect of heat, which then results in a change in preload.

I wouldn't use a plain bushing in a modern machine. Rather, a cylindrical roller bearing used at the outboard end can provide good centering without requiring any lengthwise preload consideration.

ABEC1....do you have to look hard for those? :D

mackeym, I've seen near every kind of bearing used on lathe spindles. All have limitions and advantages.

If you use angular contact bearings, keep in mind the spindle shaft will grow with heat. The farther apart, the greater the growth and increase in preload. Also, an ABEC 1 is not the most accurate, it's the least. I'd rather see an ABEC 5 or better.

You will get many recommendations on material. Personally, I like 4140/50 or 4340. Your spindle should be hardened to improve its mechanical properties.

If you elect to use bronze bushing type bearings you MUST harden and grind the bearing surface of the spindle or you will wear the spindle as well as the bushing; much more hassle to replace the spindle than the bearing.

You might also consider tapered roller bearings (Timken brand has become almost a generic description). They can be had at the low-cost-not-so-accurate level as well as the high accuracy / expensive end.

If you choose bearings by TYPE and not true calculated bearing capacity versus applied load potential, you have no clue or assurance that the bearings will live let alone perform. Somewhere on the 'Zone, one of my prior rants about bearing load life calculations will go into further detail why this is the case.

The A/C's are there to control axial displacement and it is far away from the spindle for a reason - low RADIAL capacity. The bushed bearing is/was placed next to the chuck to accept the high radial loads inherent to lathe loading at that position.

Don't want the friction of a bushing? Consider a precision cyllindrical, and specifically one with a tapered bore. These are adjustable so as to set the internal clearance. THe use of fixed clearance garden variety cyllindricals have too much internal clearance for good finish and/or tolerance control.

You can run properly opposing tapers or A/C's, providing you account for axial growth under heat so as to NOT overload the bearings if the spindle grows with heat or unload the preload if the mounting position causes the opposite situation.

IF however, you chose to take a SWAG and run what you have giving no concern about load/life calculations, you can pretty much use anything. It may or may not work and nobody who REALLY knows ANYTHING about bearings will be able to say if it will work or how long it will live.

Caveat emptor.

ok. Thanks for the excellent tips and information. I'm going to go with 2 angular contact bearings (light preload) on the left side and one or two tapered roller bearings (or 2 angular contacts) on the right....and also think about thermal expansion some more and do some calculations.

One more question. To eliminate radial play between: 1. inner races and the spindle and 2. the outer races and headstock, I was thinking about making the parts a little bit too small/big, heating them up, push them together, so they lock together when they cool down. I would heat the parts to acceptable temperatures (not too hot), and also do some calcs to see what kind of expansion to expect. One fitting would be left loose to allow the spindle to expand axially. Are there any obvious problems or concerns with this?

Interference fit (press fit) can be a real PITA. I much prefer a modular means of locking the bearings in place... usually a cover plate for the outer race and shaft collar nuts for the inners. I preload using wavy disc springs.

I'm with DZASTR on the spindle material. Among those he listed, 4140 Prehard is my material of choice.

Is it possible to turn 4140 prehard or do you have to machine the unhardened metal and then have it hardened? If it's hardened after machining, will there be much dimensional change from the hardening process?

By the questions you are asking, you're a long ways from ripe yet :D Pretty ambitious project for a beginner ;)

4140 prehardened is machinable with carbides quite readily. Drilling it for great distances is a bit more difficult due to the casual user not having the best equipment available for the job, but if you take your time, and use lots of coolant, you can drill with with HSS drills.

The bearings are usually a slight shrink fit on the spindle, and a very close 'tap in' fit in the housing. This is due to the simple fact that assembling/disassembling the contraption is a beast of a job if you do it otherwise.

If you go with two sets of angular contact or tapered rollers, one set should be constrained within a housing, but the other end should be able to float (axially lengthwise) otherwise you will still have undesirable preloading occurring due to differential heat expansion between the spindle and the housing. The spindle runs hotter than the housing in most cases.

A single tapered roller or angular contact bearing is useless. You would be better off with plain ordinary radial ball bearings on the outboard end of the spindle. If you have the bucks to spend, you can find ball bearings with lower than normal clearance, even precision grades might be available. However, the outboard bearing is just a steady basically, the front bearings do most of the work.

One question: how are you boring your headstock? Are you quite certain that you should not be starting off with the purchase of a lathe, new or used?

mackeym, None of us here who are attempting to help you have enough information about your project. Give up more specifics and we can be more helpful. When it comes to bearings, pay attention to NC Cams, he's more up on the subject than most or all of the rest of us. He does this stuff 'cause he's dedicated. I do it 'cause I'm retired and bored stiff.

Thanks for more good info. Yes, i did think about redoing a used lathe. I was going to retrofit a micromark mini lathe. However, the only thing that would be left is a headstock, so i decided on building a larger one from scratch. (i'm going to use SHS25 rails for both axis's and ballscrews with AC brushless servos). I already have the ballscrews (C5 and better preloaded), one set of SHS (the other on the way), i have the motors and drives setup and working, and two 10:1 gear reducers (Bayside Stealth model).

For the headstock, i wanted to use a 12" x 12" x 2" granite surface plate for the base (i know that a special drill bit is needed for drilling mounting holes). I would then mount two 5"x5"x5" angle plates onto the surface plate giving two vertical supports. Then buy 1" x 8" x 18" flat ground stock (low carbon) and bolt this vertically onto the angle plates. But, before the ground stock is mounted, i will mill (CNC) out an area for the bearing outer races. I think that alignment of the holes for the outer races is important, so i plan on coming up with a way to ensure the alignment (after they're mounted) is better than 1mil.

And, is this going to be some sort of ultraspeed machine?

haha, no I'm more interested in accuracy than speed (trying to get 1mil or better).

The A/C's of a P5 (ABEC 5) tolerance should be fit with a shaft tolerance of "h4" for ID's from 10 to 80mm and a "js4" for 80 to 200 mm shafts. For a P4 (ABEC 7), use "h3" and "js5", respectively.

THe same fits would suffice for tapers BUT finding ABEC 5 or 7's will NOT be easy or cheap. You might want to loosen up the tapered fit a bit to account for the sloppier tolerances you may have to live with in tapers.

The above yields a target INTERFERENCE that varies from 0 to 2 and up to 5 microns according to the size of the shaft and bearing ID.

The Fits in housings depend on whether the end has to remaing free or floats.

In your case, you could FIX the A/C end and, if using tapered rollers, set them up with as close to 0.0000" end play as possible and either DB or DF mounted but then let them float. This would eliminate any thermal growth issues from affecting the axial load on the tapers.

For a fixed end bearing (the A/C's in this case), the housings for P5 grade bearings get fit with a "JS4" while P4's get fit with a "JS3"

For a slip fit bearnig (the tapers in this case), the housings should be fit at a H4 for P5's and H3 for P4's.

These housing fits range from 0 to 5 to 10 microns LOOSE.

Any bearing handbook will explain what these metric fit tolerances mean in numeric terms.

Anything short of the above tolerances would constitute SWAG fits which may or may not fit nor work. You're quite welcome to follow or ignore the above industry standard machine tool fit recommendations. One makes their choices and takes their chances.

EDIT: For those interested in my prior rants regarding the critical need for proper bearing fitment, do a "bearing fit" search on the website. Basically, if you want the optimum performance and accuracy from a device, especially a machine tool spindle, bearing fits are of paramount importance and should not be taken lightly. END EDIT

RE-EDIT: Shaft materials can vary but need to be properly selected in order for the shaft to live/perform properly. For optimum stability and accuracy, the shaft should be turned and then heat treated and then finish ground. For rolling element bearings, PHT4140 would work if it were H&T to Hrc 28-32 minimum. I'd like it better if it were H&T to Hrc 44-48 as this would make for a much tougher, stronger part.

An alternate material for sleeve bearing/bushed shafts would be a carb and hardned steel. A garden variety yet good steel to use would be 8620. For a small price premium, get 8620 AQBQ (aircraft quality/bearing quality). The AQBQ grade is cleaner and nicer to machine.

The problem with any procured steel anymore is that mill run steel seems to be all over the place in hardness and residual stress. Hence, we find that billet tends to move around quite a bit, especially when you cut into it. Shafting is especially sensitive as it twists and bends as you start to machine it.

The fix: leave grind stock on the part. Subject the part to a 1200 deg stress relieve for 1-1.5 hours minimum prior to subsequent H/T. Cool to room temp prior to H and T.

4140/4150 is H and T to Hrc 44-48, double draw if possible.

8620 then gets 0.060" case carburized for use with sleeve bearings/bushings. Harden and double draw.

With either steel, it is always good to cryogenically stabilize and stress relieve after cryo.

Do the above and you're shaft will be darn near bulletproof, stay straight as an arrow and probably outlive you as long as you don't do something really dumb when you design/machine it.

BTW, we do custom/specialty OD and ID shaft grinding of round parts for specialy applications in case you can't find someone locally to do it. Send a paper drawing for a quote.

END RE-EDIT

Thanks again for the info. I'm working on figuring this stuff out and finding suitable bearings. l'll post questions, comments, ideas, new headstock plans, as they come.

For bearings, i found 2 options:

1.
Taper roller bearing FAG 32016X.P5 Cost ~$150 ea.
This is claimed to be a "Iso Class 5 (RBEC 5) tolerance" bearing. A P5 bearing. I would guess that this has an accuracy equivalent to an ABEC 5 bearing.(?)


2.
Angular Contact bearing SKF 7216 BEGAM Cost ~$150 ea.
This is an "explorer" bearing and has a dimensional accuracy of P6 and running accuracy of P5. These have the advantage of simple preloading: the faces are ground so that it is preloaded by just clamping two inner or outer races together.

SKF has a nice online load-life calculation. The angular contact bearings can handle a decent amount of weight. 1000lbs in both radial and axial directions at 1000rpm in iso680 oil will last 68yrs (worst case senario) of continuous running. This load-life should be fine for anything i'll be doing.


Comparison of each:
Advantages of the A/C: Easy preloading
Advantages of the taper roller: Higher load capacity. Higher dimensional tolerance/accurcy.
Spec common to both: Running accuracy of P5



Right now i'm leaning towards using two pairs of angular contact bearings because the preloading is easy and i don't need the higher load capacity or dimensional accuracy of the tapers.


Regarding Shaft and Housing Fits: Basically, to get a reliable fit and follow industry standards, the shaft must be ~2 and up to 5 microns larger than the bearing inner diameter for both bearing pairs. The housing fit for one pair should be 0 to 5 (and up to 10) microns loose. The housing fit for the other pair should have ~2 to 5 microns of interferance. I think it may be good to have the housing interference fit on the right side (adjacent to chuck) so thermal expansion doesn't affect axial accuracy. and the loose housing fit on the left.


For shaft materials, i think i'm going to go with 4140 PHT and possibly have it rehardened to higher hardnesses. Not yet sure how i'm going to bore it (or get the housing machined to the required accuracy). I'll think about those things next.

I may post a picture of the proposed design sometime this week.

For bearings, i found 2 options:

1.
Taper roller bearing FAG 32016X.P5 Cost ~$150 ea.
This is claimed to be a "Iso Class 5 (RBEC 5) tolerance" bearing. A P5 bearing. I would guess that this has an accuracy equivalent to an ABEC 5 bearing.(?)


2.
Angular Contact bearing SKF 7216 BEGAM Cost ~$150 ea.
This is an "explorer" bearing and has a dimensional accuracy of P6 and running accuracy of P5. These have the advantage of simple preloading: the faces are ground so that it is preloaded by just clamping two inner or outer races together.

SKF has a nice online load-life calculation. The angular contact bearings can handle a decent amount of weight. 1000lbs in both radial and axial directions at 1000rpm in iso680 oil will last 68yrs (worst case senario) of continuous running. This load-life should be fine for anything i'll be doing.


Comparison of each:
Advantages of the A/C: Easy preloading
Advantages of the taper roller: Higher load capacity. Higher dimensional tolerance/accurcy.
Spec common to both: Running accuracy of P5



Right now i'm leaning towards using two pairs of angular contact bearings because the preloading is easy and i don't need the higher load capacity or dimensional accuracy of the tapers.


Regarding Shaft and Housing Fits: Basically, to get a reliable fit and follow industry standards, the shaft must be ~2 and up to 5 microns larger than the bearing inner diameter for both bearing pairs. The housing fit for one pair should be 0 to 5 (and up to 10) microns loose. The housing fit for the other pair should have ~2 to 5 microns of interferance. I think it may be good to have the housing interference fit on the right side (adjacent to chuck) so thermal expansion doesn't affect axial accuracy. and the loose housing fit on the left.


For shaft materials, i think i'm going to go with 4140 PHT and possibly have it rehardened to higher hardnesses. Not yet sure how i'm going to bore it (or get the housing machined to the required accuracy). I'll think about those things next.

I may post a picture of the proposed design sometime this week.


Well, Don´t know what size you are looking for but I have 2 pairs of 7013 and 2 pairs of 7012 superprecisionbearings 4 sale.

FYI: these are 1350$/pair here.

I want 200$ /pair.

http://www.emersonbearing.com/SpecSheets/SuperPrecison.pdf

News flash: ABEC (ball) and RBEC (roller bearing) accuracy specs are for 2 entirely different types of bearings. Until/unless you compare the specs, line by line, spec by sped, you should not assume that the absolute magnitude of each class is/are are equivalent. Details why a bit later

The most important spec you want/need to worry about is concentricity of the inner ring rotation, one to the other and/or radial runout of the rotating ring. THis is why eccentricity highpoints have to be marked and matched when mounting precision bearings in duplex fashion.

I know of a particularly embarrasing situation where two bearings of different sizes but same accuracy class and differeing types (roller and ball) were coaxially mounted on a shaft. This was done to gain the relative merits of each with regard to axial and radial capacity in a very tight package space.

Imagine the surprise on the engineer and customer's faces when after a few revolutions by hand and under NO load, the shaft fitted with this concoction would sieze up totatlly. When they took the thing apart, NOTHING out of the ordinary was noted, no scuffing, no scoring, NOTHING. There was absolutely NO damage.

So they reinstalled the bearings and started rotating it again - they did this a number of times in fact. Same thing each and every time: somewhere between 8 and 12 revs after you started rotating the shaft, the thing would sieze.

The problem was a mystery until you looked at radial runout of inner rings between the two bearings. Although they were of the same accuracy ranged, the bearing SIZES allowed them to have different MAGNITUDES of radial runouts.

Result: after a number of revolutions, the axis of rotation estabished by the one bearing "shifted" due to allowable runout. Whe the same thing happened to the axis of revolution of the other bearing - siezure..

Basically, after a number of revolutions, the "axis shifts" were to the point that they were so much in conflict that the internal clearance was consumed and the shaft siezed - it had no choice but to sieze.

Nothing and I do mean NOTHING but absolutely perfectly matched and assembled bearings would fix the situation. Since this was a high volume automotive application, where speed of assembly is critical and hand selection/fits all but impossible, the design was shipped "back to the drawing board for further study".

Don't forget, by the way, about housing bore concentricity to each/one another.

The higher the bearing accuracy the closer the housings on either end need to be with regard to coaxiality. If they are not, the shaft will have a tendency to sieze as the shaft axis orbits about as the ball and/or roller runouts combine with the raceway runouts.

BTW, the coaxiality factor becomes more and more critical as you both increase the accuracy of the bearings AND as you preload them. The higher accuracy bearings have less self centering/clearancing "slop potential" and the preload takes out the radial and/or axial clearance. THus, the bearings are even less tolerant of "off" conditions.

WHen you are dealing with fits in the single digit mircron level and runouts at or near the same level, you really have to dot your I's and cross your T's when you start to size and fit bearings.

NOTE RE: "capacity" = just because the bearing has 1000 lb load CAPACITY the does not mean you can apply that much load. Most fatigue life calcs ultimately assume that only a small fraction of the rated capacity will or can be applied for a 100% duty cycle. Apply the rated load, even for a short time, and the bearing life can/will drop in a precipitious fashion.

Again, I can cite stories where an engineer did a "catalog load rating review" life calc, selected the bearing and the client used it. Imagine the surprise when the bearing failed almost immediately.

Just like in the useage of accuracy specs, load life calcs need to be done properly and likewise interpretted in order to realize the full life and performance potential of any bearing.

gotis,

Do you have any similar but larger bearings? 16's (80mm ID)

gotis,

Do you have any similar but larger bearings? 16's (80mm ID)

No, sorry i don´t, these where for a cnc lathe project I was working on but I found a complete spindle cartridge instead.

By the way, these are paired bearings, I can only sell them in pairs

mackeym, My last spindle design used a pair of A/C's similar to your proposed at the chuck end. (use good seals) and a floating deep groove radial and nut just ahead of a belt drive pulley. The rear cover had another radial bearing in it help support the tail end of the shaft against belt tension. (Sort of a belt/suspenders arrangement) Of course, you must still observe the basics (see NC's remarks). I don't think I would hold a second pair of A/C's on the right end. If you want the extra capacity, why not mount them as tandem duplexed pairs behind the chuck?

...but I found a complete spindle cartridge instead.

gotis, was this an eBay find or something regularly available? If the latter, what kind of accuracy/speed/price range?

Thanks,

Randy

Richard,
I was thinking about using A/C's near the pulley end to minimize radial moment play. If the left side has some play, the right A/C's may act as a pivot point for radial play in the work piece. This moment play would be more noticable the longer the work piece. To minimize it, all radial plays on the left puley end need to be minimized. If i use A/C's on the pulley end, then this moment play would be caused only by the lose housing fit which has to be presant. But, maybe i should choose a less than normal internal clearance bearing instead of two A/C's

SInce the A/C's are supplying the axial thrust resistance, you should FIX these into the housing. As in FIX them AXIALLY with a clamp plate - you need to do that to affect any preload. Also, FIX them in the housing RADIALLY with essentiall the INTERFERENCE fit as prescribed in my prior post.

This could be done at the pully side if the A/C's have a lower radial capacity than what you need for the chuck but there might be a better way - please read on..

When it comes to the chuck side, this side should be SLIP fit in the housing (crudly put as SFNS - slip fit no slop) and what ammounts to LINE ON LINE with the shaft. Thus, when the shaft grows/shrinks with heat, the bearing is free to move axially in the housing as need be. The trick here is how to create such a fit and to coincidentally establish the proper clearance in the bearings - especially if you're using tapers which do have essentially an ability to be set at clearance or even a slight preload.

However, when you look at any number of spindle cartridges, they do things differently. They tend to run a tapered bore ROLLER bearing on the chuck side. These self align axially and have a tremendous amount of axial capacity. The tapered bore allows you to adust the clearance to nearly nothing which essentially provides loads all the rolling elements simultaneously - it also removes most if not all the radial play which will effectivey match the radial clearance of the roller bearing to what's going on in the preloaded A/C's. THus, it is perhaps better to FIX the chuck side for optimum runout and rigidity at the chuck.

Another option is to mount a QUAD set of preloaded A/C's to the spindle side. FIX this end in the housing and then let the sprocket side float in the housing. Sometimes, they even mount the spindle side in a "carrier" which slip fits in the housing. THe carrier is then spring preloaded with about 1%-2% of the radial capacity rating of the bearing. This provides probably the truest running, slop free spindle side bearing you could concieve.

QUADS are pricey but they do have both radial and axial capacity - something to consider.

If you go with tapers, you end up having good, reasonably priced very high capacity bearings. The only issue is adjusting them for optimum "axial clearance" - forget checking radial clearance as that's pretty much impossible to do in the field.

Tapered rollers are one of the few bearings that can function quite well with essentially NO clearance and/or even high axial preload. In fact, you actually preload pinion bearings quite high to maintain pinion position and to resist axial thrust induced unloading of the bearing preload.

For a lathe, a rule of thumb would be to set up the tapers with 0.0000" axial clearance. This is done by reading axial clearancee with a dial indictor while you first rotate the shaft CW and stop when the indicator stops moving. This is now reset to ZERO.

You then start rotating the shaft CCW and the axial clearance is the value when the indicator stops moving in the other direction. You simply keep reducing axial clearance and rechecking until the indicator stays at 0.0000" when rotated in both directions. At that point, you have NO axial play in the taper, essentially NO radial play potential and almost the absolutely the stiffest bearing you can come up with on the cheap. The runout accuracy of the bearing will be established by the RBEC specs you got when you bought the bearings.

To summarize, FIX the chuck side bearings - FLOAT the spindle side.

Choose bearings based up on need and capacity, not purely on availability. Design in what you need, then get it. Designing something to use whatever you can get hold of could yield anything but what you want/expect.

Prototrains.
Sorry I got it cheap from a toolmaker who was changing from weldon to hsk spindles, they where running 24h 5days a week at 2500rpm and usually don´t need to change bearings more than every second year, (SKF cartridges)

Machine tool spindle design is based on a few simple concepts. Get too far away from any of these and you will have problems. The granite with the angle plates with the flat stock is well off the beaten path. I admire your creativity but,the fact you said milling the bores is your plan tells me and anyone else who has learned this stuff the hard way,you need to do some reading.

Regarding bearings, i found a matched pair of ABEC 7 A/C's (new and cheap) for the chuck end of the headstock. Regarding loading, i plan on staying below the fatigue load limit (and no where near the basic load ratings). For the pulley side, i'm going to use some type of low clearance bearing (which i havn't found yet). NC Cams, regarding your concerns with the 1000lb load on some previous bearings, that load was near the fatigue limit. Correct me if you know otherwise, but as long as you stay below this limit, the bearings should last almost forever (assuming they're lubricated and kept clean). If this is wrong, then the bearing manufacturers incorrectly designate fatigue load limits which is doubtful. You guys probably know this but i'm not talking about the basic load ratings; i'm talking about the fatigue load limit.

Thanks pepo. I'll look more into headstock boring.

In order to calculate the life of a bearing, you have to use a load which, if applied 100% of the time and does NOT fluctuate, ever, will cause the part to fatigue. It is important to realize that this is for the CALCULATED fatigue life.

Can, will, should you load a bearing at/to its fatigue load in a one shot basis and could/should/will it live? NO to all the above.

I've seen cases were clients have loaded bearings to high percentages of the fatiugue load and saw the bearings fail almost immediately. Yet, I've also seen where a client has applied a judicious and disciplined break in cycle, a slowly ramped up load and speed application rate and they ran at or near the fatigue limit for longer than the calcs would suggest were ever possible. Why's that?

Because once you do a load life calc, there are then ARBITRARY (and somewhat capricious) rating/derating factors that need to be appled. The take thinks like lube life, contamination, alignment and other variables that can/will/do affect bearing life in to consideration. Sometimes these factors are published but more oftern than not they are in the form of "sage wisdom" that is a tuning tool of the bearing company's application engineering dept.

This is why the bearing suppliers pretty much ALL have the "the customer sould consult our technical dept for specific information and recommendations".

Some of the worst application engineers are those who do their load life calcs via a SWAG of the catalog and mentall apply a simple proportion of the fatigue load to a/the bearing and pronounce it satisfactory. Taking the same ratings and applying the ENTIRE prescribed load/life calc process, I've taken the same figures and same loads and applied the published rating/derating factors and found that the previous "no problem" bearing calc to be quite deficient.

Essentially, the bearings that shouldn't have failed according to the client's calculations lived longer than they should have when calc'd via the factory prescribed method.

IF you have used the ENTIRE load/life calc process that's prescribed by the bearing maker for/of the bearings you choose and feel that you've properly done so, I can't argue with your conclusions. However, if you did a catalog scan and simple compared some arbitrary loads you think you'll be seeing to those published in the catalog, well, all I can say is that my prior experience with this method has not been a glowing history of satisfied, happy self taught application engineers.

Caveat emptor.

ok, thanks NC Cams for more good info. So basically, you have to go through the entire load-life calc process with all variables considered (which is what i did). My interpretation of the last post is that the bearing generally lives longer than what this "total calc" gives (although there are most likely many exceptions and instances where this isn't the case). If one wants any higher of a confidence level, then you have to talk to the company's technical department.

"....If one wants any higher of a confidence level, then you have to talk to the company's technical department..."

In a word, YES, that pretty much sums it up. In reality, the load life involves periods of low load and occasional high load. The calc of a real life load profile tends to be tedious and troublesome. Why? Because there are always assumptions and guesstimations than can affect the load life calc one way or another.

Consultation with a tech dept can ofter uncover a guy who's had experience with an application that can help "reality-ify" the load life calc with an emprically derived modification factor. Or, you can find a slug and not get any real benefit for your call

If you've done a full blown lift calc, use the proper tolerances and fits, you should have as good or perhaps better than the predicted life. If not, then the opposite should also be true.

I wish you well with your project and do hope that you did your calcs well and properly..