Shaft material opinions?

Here's some thinking out of the box type stuff that has been used for shafts in much LOWER power applications. Don't know if it can be scaled up to your power levels. Spring steel ie. piano or music wire used by Rick Willoughby in human power applications, 10mm up to 0.5 hp. I have use fiberglass shafts up to about 100 watts. Both are mostly indestructible because they give and realign upon severe impacts. Spiral wound or braided carbon fiber is much stronger than fiberglass and might take the impacts. Bushings are glued to the shaft and then ball bearings to mate with the bushings. FWIW.

Porta

 

Must say I share Frosty's opinion on your coupling design; set screws plus key are crap engineering, and they were so a hundred years ago as well! At that time, good engineering had been a tapered end with double nuts. Today, the prefabricated clamping units do the same thing easier and better. Attached you will find three varieties used here; I am sure you will find those all over the globe. I don't remember having milled a key recess for many years now! Shaft keys are only acceptable if components have to sit at a specific angle on the shaft!

The tabular texts should be selfexplaining, suffice to say that "daNm" is torque in Nm/10, ie ~kpm. The first page shows the necessary outer hubdia for various hub materials.

Now to the shaft material. With the engine power, rpm's and shaft dia given, only the Duplex steel 1.4462 as mentioned by Apex1, has the strength required for full power! All the others will fail due to either overload or fatigue. In addition, the bronze shaft has only half the modulus of elasticity, compared to steel. As the shaft length is close to the allowed limit for one bearing, the soft material will thus deflect under rotating load, causing premature failure due to the combined effects of torque+radial deflection+compression.

In order to save your props when shoveling rocks, use a short brass key as a torque limiter. Fit propeller to taper, use a closed hub-nut plus shims to press the prop ~1 mm onto the taper (with hub-nut bottoming on shaft end).

 

Sorry Beckmo, but I disagree that the stress is that high in the shaft, or that he needs a material such as a duplex steel to work if he has the right coupling and prop.

At 105 hp and with the engine running at 3200 rpm, the shaft shear stress is 10.5 ksi in a 1 inch shaft. Even the 316 stainless can handle that assuming a 42ksi yield strength and a .577 shear factor (max stress to prevent yield is going to be 24 ksi). You need to avoid any big stress concentrations, which is why the a tapered coupling is a good idea, and I probably would like a margin for any stress concentration, but a 2.3 factor of safety is likely ok. It isn't a huge safety factor, and could have a lot to do with why the shaft failed in under the pinch screws, but for a proper tapered coupling it will work fine for a long long time.

Looking at the bending stress in the shaft at the prop end, and taking your numbers for thrust at 232 kg of force fwd and 91 aft, and converting that into a moment of 323 kg f (712 lb force) at a 5 inch radius (avg 10 inch dia of action on a 12 inch prop dia) I get a total moment on the shaft due to prop thrust being uneven is about 1615 inch pounds. This results in a stress of about 16.5 ksi in bending, or about 1/3 of the yield strength of 316. I don't have a modified Goodman diagram for 316, but without a significant stress riser the shaft isn't going to break as a result of high cylce fatigue with that load in it.

I agree that the margins aren't huge, and I'd like them to be bigger, but it isn't going to twist right off in short order either. The fact that he has successfully run the 316 shaft for a good long while, even with the nasty prop offset loads indicates that the system will work if he doesn't solve his prop issues. If he does fix the prop, the shaft cyclic bending stress only goes down and things even look better.

 

Yes, you are correct there, as long as you check on the "intact" part of the shaft! Admitting a slight "constructive laziness", and a bias towards design for commercial duty, I used the shaft supplier's charts for dimensioning, and came out with my conclusion. This is based on the critical area at the propeller end, where area reduction due to taper and due to key notch combine to a high, asymmetrical stress concentration. Please note that I wrote "....strength for full power...".

Note that the "force" figures from the other thread are for the infinitesimal area strip dA at 0.7*r. You have to integrate this product over the whole blade area in order to use it in another context! The total propeller thrust is something like 3500 N, still with the asymmetric pressure distribution indicated by the ratio given in the "rake" thread. This increases the shaft stress considerably. Also note that the force asymmetry plus the blade pulsations result in a pulsating load on the shaft.

Another factor is the shaft length/dia ratio. With a fixed coupling, the minimum length between supports should be 40d, ie 1000 mm (dNV). This shaft is 832 mm, and should thus have a flexible coupling. There are also limits for maximum length between supports (Lloyds give a max length of about 600 mm for a 25 mm shaft whirling at 3200 rpm!). Unless this rotating mass is carefully enclosed; safety first!

 

I wouldn't worry too much about shaft length. The overall length is about 32.75 inches, the length between supports isn't any higher than 28 inches, which gives you a first mode critical speed of over 6,000 rpm. In addition there is some centrifugal stiffening from the prop, so you aren't going to see a shaft whip problem until you are probably over 7,000 rpm in the real world. Even if the shaft length between supports was 32.5 inches the first mode would be at 4500 rpm, which is a 45% margin on critical speed, which is plenty.

The shaft bending stress then is a good bit higher, which may indeed be the limiting factor..

 

"~1 mm onto the taper" - One mm from where? I like these couplings. I used these on yachts - http://www.skf.com/portal/skf/home/products?maincatalogue=1&lang=en&newlink=12_1_1

 

Yellowjacket, I'm buying your engineering arguments on shaft length hands down. What I was trying to say was that when you give advice to a client, there are two important additional issues:

A) There may be mandatory design rules covering the subject.
B) Supplyers limitations for their equipment.

If your design, or advice, does not adhere to the possible limitations from those sources, any faults occurring to, or as a result of, the equipment (or, in fact, connected equipment as well) is YOUR responsibility, no matter how careless your client has been! You are the one who has to prove that your design is ok if you don't have a written acceptance either from A and/or B.

A "non-responsibility" chart from your client will not save your ***, neither legally nor in terms of professional reputation.

 

"...A "non-responsibility" chart from your client will not save your ***, neither legally nor in terms of professional reputation..."

Yup, that about sums it up. Since we don't know each other from a bar of soap, nor if all the information given is correct, or being done as suggested. So, least we can do is, well, summed up by A) and B). Fully concur.

 

Beckmo and Ad Hock

Points well taken....