Edward Thompson: Wartime C.M.E. Discussion

The GW 47s had sideplay and inclined planes on the rear axleboxes to provide some lateral control. They did have a problem with stability ('tend to nose about a bit' - Cook) at speed and were limited to 60mph. Apparently when the running department asked for some more Collett elected to build some extra Castles instead as being more versatile but more expensive.

 

Many thanks for that. However it does rather beg another question, one best asked on a more appropriate thread.

 

Yes, but what you don't seem to have taken into account, as I understand it, is that all these fractures took place directly behind the driving wheels. The other end of that axle part fitted into the inside cylinder crank web also had a presumably similar key and keyway. So the question remains as to why did the axles fracture at the 9-5/8" diameter behind the wheel rather than at the 8-1/4" diameter adjacent to the inside crank web?

 

To the Mod - thanks for sorting that one out.

 

A shaft that is 8.25 inch one end and 9.62 at the other that transmits pure torque will have specific stresses that are 1.58 times higher at the small diameter.
Pure torsion overload cannot have been the only reason for P2 trouble I think.

 

I am not seeing keyways in the Crank Web as suggested.

That to me suggests torsion failure. Although the shaft is smaller at the web, there is a degree of twist along the shaft which is at its maximum at the wheel. If you were to imagine the wheel as fixed, and the centre piston applying thrust through the crank, there would be a twist in the shaft. That twist is at its most acute at the furthest axial distance from the application of the force, due to elasticity in the shaft. That position is immediately inbound of the wheel. The suggestion that fatigue fracture is also playing a part in this failure is vindicated further in that the new design has stress relief groove in the axle, immediately inboard of the wheel set.

If there is significant torsional twist, and there is a significant stress raiser in that keyway in the original design, then the problem is almost certainly one of fast fatigue fracture. Fast as in below 10^7 cycles IIRC. The image of the failure at Stonehaven certainly supports that theory.

I am not seeing keyways in the Crank Web as suggested.

 

I don't follow that (the bold bit) at all.

 

If I had been asked how to make wheel-crankshaft connection ,a pressfit with key solution would be the last option.
The best connection has always been the thermal shrink fit like between stub and web shown and no key.They are not easy to dismantle and if that is needed SKF oil pressure method was used on quite some swedish locomotives with roller bearing crakshafts.
The mating parts are made with a very narrow taper and two O-rings.When parts are not able to go further oil is injected beween the two O-rings and parts are put in final possition and Oil pressure released.
I have seen ship diesel crankshafts made that way and they did not do uggerley things.

 

Pictures as they say. Its not as straightforwards as that because the shaft is constrained at both ends, and the wheel itself does move, it would take a little modelling to show the true stress distribution.

What is really going on is that you have partial and highly variable constaining as the machine goes through its motions. Although the P2 is sure footed, it can still be made to slip under the right conditions. With perfect rail conditions, stresses in that shaft will be considerably higher than a wet autumn day.

What we really need to nail this whole topic is the stress report. That will show the loading scenarios used in modelling the new shaft. It would also show the S-N curves which would really enlighten us as to how the shaft is designed.

Suffice to say, the new key way is ramped, and the old one is a sharp corner. Lessons have very much been learned.

 

Does anyone have any contacts in the Trust, to ask if that can be released, to offer insight into the historical failures?

It would certainly be ironic if the P2s' major goal (better traction) turned to be behind one of their greater failings (broken axles).

Noel

 

If the shaft is uniform along its length then it makes sense that the angle of twist is proportional to the distance along. But that is the angle relative to the fixed end. The degree of twisting dΦ/dx is constant.

 

Quick update. We are now at the exciting point of getting the manuscript ready for final sign off and printing.

I have to do one final sense check when the book's layout is "complete" and then my publisher will sign it all off.

So we are getting closer and closer. Really pleased with it all. My publisher has been very supportive of the book.

 

Looking forward to it very much.

 

+1

Richard.

 

+2

 

I've been mulling over this over the last couple of weeks, so making a posting isn't helped with this subject being spread over two topics. Perhaps some concentration could be done by the Mods to transfer this subject into just one of the topics would help.

Starting with the outside cylinders, the power from them goes to the crankpin which in turn divides the power via the coupling rods to the other coupled axles. Thus it could be argued that only 25% of that power is applied to the crank axle. I accept that momentarily it could be a bit more than this while the clearances between crankpins and coupling rod bushes are taken up, possibly by a small amount of wheel slip of the crank axle.

Turning to the inside cylinder, the power from that is divided 50% each between the two crank webs. This suggests that more power is applied to components of the inside cylinder crank than the outside crank. From a torsional point of view and looking at the inside crank webs and the axle halves shrunk into them at 9" diameter and 5" long and without a key, on the face of it, there is more risk of failure in that area with the axle halves being only 8-1/4" diameter adjacent to that point.

As there are no recorded failures in that area it could be said that the inside crank assembly was certainly adequate for its purpose, yet the axle failures occurred at a 9-5/8" diameter behind the wheel, where with a lesser load it could be said to have been over-engineered. This is why I feel that there is a bit more to this than just torsional stress.

 

Sure, but what is the more that you feel is present?

 

That a sideways revolving and bending motion of the wheels on curves is a factor that has not been taken into consideration.

 

On the WD 2-10-0 there was total half an inch side play on first and last driver.
Let us assume one inch on Ps and nil on the crank shaft.Enter a very mild curve at speed and first driver will not be part of guiding.
The Gresley patent double swing link pony was designed before Cox and Stanier had been to India and discovered that british pony and boggie trucks were miles to soft in transverse stiffness.
P2 Pony truck much to weak and next to no support against angular movent and first driver flanges no touching outher rail means thank crank has to bring all the tranverse force.As it is close to mass center of locomotive it is a horrible lot of force and with a keyway at the most stressed point it is no wonder cranks broke-.
The DeGlenh,Churchward ,Cox Stannier system with side bearings high friction ,high damping material and stiff transverse srings was the way to go and Thompson did it.
Giesel Gieslingen describes how he calculated a suitable initial side resisting force on 4 tons for the 2-8-4 with 2 meter drivers where contemporary british value was a ton or less.
Gressley had it coming.

 

But as has been pointed out, the short axle sections outboard of the axle support bearings (I'm drawing a blank on their name - axle boxes, I think is the term) and inboard of the wheel also have the very considerable bending stress of supporting the weight of the loco, which is almost certainly why the diameter of that section of the axle is larger than the diameter of the inboard section (next to the crank webs).

Noel