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Hi s.sculella,
The lead error is ± .006 inch/ft.
Repeatability is a system feature and not part of any screw or nut spec. It is a function of the load, bearings, mountings, coupling, gearbox or drive train, motor, controller, feedback device, operating conditions, etc….in addition to the screw and nut.
Typically users build the positioning system and then measure the repeatability at intervals during life testing conducted at rated loads and operating conditions.
Thank you for using our products, and let us know if we can provide more information.
Normally steppers have no thrust capability so direct coupling is rarely done. You can try it and see what happens.
Most installations isolate the thrust with separately mounted thrust or angular contact bearings and then couple with a flexible coupling to allow for some misalignment.
For the material we recommend acetal plastic (Delrin or equiv.). The internal threads should be ½” maximum length.
We don’t make couplings. Most users couple directly to the stepper motor output shaft with commercially available, helical cut aluminum body couplings with set screws which work well for limited loads. I would use ¼” output shafting on the stepper and ¼” bore couplings and you won’t have to do any machining.
Any method to prevent rotation will work, mill flat(s), a keyway, pin(s), etc…
The efficiency of this size is 66%, see column “H” at this webpage: http://www.roton.com/Engineering_Data.aspx?line=Hi-Lead
The efficiency is the same for steel and stainless steel. Efficiency is a function of lead angle, flank angle and coefficient of friction only. Here’s a link to the formula webpage (see # 10): http://www.roton.com/formulas.aspx
The preloaded nut is de-rated because the preload is very light and only effective for small loads. There is also a limit due to the flange mounting. We do have users who load these beyond the catalog rating of 25 lbs. without any problems but each application has be to tested to make sure it meets the service life and performance requirements.
You can use two sleeve nuts and make a more robust anti-backlash mechanism if PN 19664 proves to be inadequate.
In simple terms, the ANSI B1.5 ACME screw standard applies to single start screws only. Once the lead angle of the thread exceeds 5° as with 2, 3, 4 and more starts, some interpretation of the normal flank angle and internal and external thread limits of size must be made by the screw and nut toolmakers and this varies widely by manufacturer. It is always risky to procure the nuts and screws or tooling therefore from different manufacturers.
Can you send us a sample of the screw and nut(s) you are purchasing?
Can you send us drawings of the screw and nut?
What is the overall length of the leadscrew(s)?
What is your yearly quantity of leadscrews and nuts?
It is unlikely that you will have success with hand tapping these nuts. The lead angle is too high and the parts will tend to ream rather than self-feed. We do this type work on large stroke, specially built CNC leadscrew tapping machines and you would need something similar that can deliver very high torque at very low RPM with the spindle synchronized to the tap slide. Do you have a machine similar to this on which to do the tapping operation?
Since these type products are what we specialize in, I am sure we can offer you something once we receive the above and have had the chance to do an engineering review. For example, we have a 5/8 X .500 size in our OEM HiLead line which may work for you. You would get 25% faster linear travel for the same RPM or you could decrease the RPM by 20% and produce the same linear speed. Generally, these are easy accommodations to make.
We look forward to hearing from you and working with you on this project.
Normally steppers have no thrust capability so direct coupling is rarely done. You can try it and see what happens. Most installations isolate the thrust with separately mounted thrust or angular contact bearings and then couple with a flexible coupling to allow for some misalignment.(Steve Lochmoeller)
A tap would cost $286. For the material we recommend acetal plastic (Delrin or equiv.). The internal threads should be ½” maximum length. We don’t make couplings. Most users couple directly to the stepper motor output shaft with commercially available, helical cut aluminum body couplings with set screws which work well for limited loads. I would use ¼” output shafting on the stepper and ¼” bore couplings and you won’t have to do any machining. Any method to prevent rotation will work, mill flat(s), a keyway, pin(s), etc… (Steve Lochmoeller)
±.009 inch/ft. These are not ground thread ball screws and are not made to any ANSI precision designations.(Steve Lochmoeller)
Sorry for the delay in getting this to you but I wanted to do some additional calculations and add more detail along with some commentary which I hope is helpful.
Here are the results for the 3 1/2 – 2 size:
Velocity Factor = .917
Pressure Factor = .0396
@ 57,000 lbs., 30 RPM:
PV = 62,000 PSIFPM
Service life = 1.68 million inches (theo.)
@ 7,000 lbs., 109 RPM
PV = 27,700 PSIFPM
Service Life = 13.7 million inches (theo.)
And, here are results for 3 – 2 size:
Velocity Factor = .787
Pressure Factor = .0643
@57,000 lbs., 30 RPM
PV = 86,500 PSIFPM
Service Life = 1.2 million inches (theo.)
@7,000 lbs., 97 RPM
PV = 34,100 PSIFPM
Service Life = 9.8 million inches (theo.)
The Factors are used as follows:
Multiplying the RPM * Velocity Factor = Surface Velocity, V (FPM)
Multiplying the Load in lbs. * Pressure Factor = Unit Surface Load, P (PSI)
P * V = PV Factor (PSIFPM)
For this bronze material, maximums are: 100 FPM, 8,000 PSI and 125,000 PSIFPM.
The RPMs used in the above results are not what you requested but based upon the above maximum.
As to wear, we use the Archard theory of rubbing surfaces which states that the volumetric loss of material is proportional* to the service time multiplied by the PV, or W =KPVT.
W = Volume of material sacrificed to wear.
K = a fixed wear rate and is a function of mating materials (the slope of the wear versus time plot)
PV = Pressure Velocity Factor of the mating surfaces (see above)
T = time
*(Testing reveals that the actual wear versus time plot is shaped like a toe down hockey stick, a very steep slope at the beginning of testing, indicative of initial wear-in, then mostly proportional on a much softer slope for the long term, steady state wear period. The Archard technique ignores the steep portion and uses the steady state slope for the wear rate. In practice, when life testing, do not extrapolate the wear rate until the rate drops to its long term slope.)
This theory is just that and in practice provides a guideline which can be used as we have used it here, to get reasonable comparatives. The K’s we used were generated in our test lab which has ideal axial loading, a clean environment and excellent lubrication. We monitored torque and heat and restricted these to very safe limits using cooling and frequent relubrication. Wide variance in life tests was still exhibited. Nevertheless, we feel these are the best guesses we can make until a better method is discovered. If you want to get a copy of Archard’s work and slog thru it all, I would recommend a full weekend with a full coffee pot with no breaks for watching football.
One more caveat in addition to those mentioned for wear life, you should use a “safety nut” whereby a steel nut is mounted in tandem with the bronze nut such that if the bronze nut fails, the steel nut will prevent the load from dropping in the event of the bronze internal threads stripping. The wear can be monitored by measuring the distance between the follower safety nut and the load bearing bronze nut. When this exceeds a threshold amount, the system should issue a warning or simply shut off. This is situation was exquisitely demonstrated by the Alaskan Airlines plane that crashed near San Diego some years ago when the elevator nut stripped out. Lack of lubrication and inadequate wear monitoring was the problem as determined from the parts salvaged from the sea bottom. (A contributing factor was the use of an aft elevator pivot such that a leadscrew/nut failure caused the elevator to jam full down or full up rendering the aircraft unflyable. Had the pivot been forward the elevator would have followed benignly in the airstream and the plane could have been flown with aileron control alone. When I proposed this to the Boeing legal defense team my name was quickly removed from the prospective expert witness list.
One more tip from an engineer having practiced the trade for some 40+ years: Build a test unit of the articulating modules of your design as quickly as possible and get some operational and wear life feedback before you start constructing the first unit which you said has to work as built. Your marketing and management team will squeeze you on the budget for this but we see so many applications constructed that have to be redesigned because the most rudimentary testing was not conducted. We have an expression here that says “One test is worth a thousand expert opinions”. That number may be too low. If you want to chew the fat some time, ask me about our experience with springs and SolidWorks simulation. (Steve Lochmoeller)All PV’s are within range. Theoretical service lives in millions of inches are:
9.8
1.8
1.2
Respectively for the 3 data sets below. This data is only a guide and we encourage you to review this webpage and the comments on wear life: http://www.roton.com/application_engineering.aspx
For info on the 3-2 Screw size (PN 60522) with standard Bronze Threaded Nose Nut (PN 91740) used for the above calculations, view this webpage: http://www.roton.com/Mating_Components.aspx?family=7060522