Member postings for Kiwi Bloke

er, 'scuse my scepticism, but I think that all the article really says is that the methodology is not standardised, and the figures are therefore essentially meaningless. It's specmanship. Remember the spec war around hi-fi amplifier 'power' ratings? RMS, peak, 'music', 'average', at what level of distortion, etc., etc.? The most important parameter relating to your breaker is that you're satisfied with it...

I find that a 12lb sledge, swung with truly evil intent, is an effective demolition tool. I used to swing a 14lb sledge. But now I'm older... The energy delivered is proportional to the velocity squared, so a considerably faster-moving, but slightly lighter walloper is better - at least for this old bloke!

Perhaps Geely, like Toyota, think the future is in hydrogen IC.

Bearing (?) in mind the wall thinness, and the hoop stress resulting from a press fit, might a Loctite product be safer?

Food calories only really mean something to dieticians. Conversion factor: 1 calorie = $ n

The waters are getting muddied by imprecise terminology. And we are getting into deeper waters than I know how to manage. Whilst 'impact energy' may be the energy required to break a test piece (as in the Charpy test), any impact upon a material will deliver energy to the material. It may, or may not result in macroscopic damage. Cracks, which lead to wholesale failure (one hopes - in this application) require energy to start, and energy to propagate. (Look up Griffiths, energy and cracks). On this machine's spec. sheet, 'impact energy' (if it means anything at all - spec. sheets can be fanciful, especially if from the Orient...) must mean the energy that each blow delivers.

Another problem is that potential energy is often quoted in units like 'foot pounds', which is dimensionally incorrect (distant memories of Dimension Theory stirred?). Gravitational potential energy = m X g X h, if you remember, so, the inclusion of the acceleration term makes the dimensions correct, as they are for kinetic energy, = 1/2 X m X v X v (sorry, don't know how to do superscripts here...). A Joule = 1 Nm, not any number of kg.m, so is dimensionally consistent, regardless of the type of energy.

However, this doesn't really tell you whether it's going to be an effective breaker, because energy delivered per area is what matters. So a sharp-pointed chisel will inflict different damage to, say a flat, blunt 'chisel' with an end area of a few square whatevers. The system is so full of unknown variables that I think the 'impact energy' figure isn't helpful.

OK, having stirred the muddied waters, I'll slink away.

Presumably the 'impact energy' is the kinetic energy of each blow of the internal hammer, delivered to the chisel. The mass (inertia) of the chisel will have an effect on the performance, but the mass of the machine itself, being huge in comparison, isn't really important in practice. How this compares to a bloke + lump hammer system depends on the bloke and the hammer, but your breaker will go on for longer, without having to stop for tea, and will achieve its results by thousands of light strikes, which will cause less collateral damage (usually...).

If beer hasn't addled my maths, 50J is about one quarter of the muzzle energy of a typical .22 round. So, if the breaker doesn't do it, shoot it - lots of times...

On reflection, I don't suppose this helps...

Edited By Kiwi Bloke on 12/07/2023 08:41:55

Edited By Kiwi Bloke on 12/07/2023 08:57:43

Pah!

Oh Bother! I'd just shut down the computer, when I realised I'd misunderstood Graham Meek's post. My misunderstanding is frequent - according to my dear wife... Anyway, it's the morning here, and I haven't had coffee yet, so you can't expect much sense yet.

Graham's reference to Emco's backlash adjustment presumably includes backlash in the nut, which must, of course be lubricated reliably, requiring space for the lubricant. In my earlier post, I was confining comments about axial float to the leadscrew support bearings only. Here, it seems reasonable to aim for zero end-float - easily achievable with rolling element bearings, but not simple with plain bearings.

Apologies for the confusion.

Please excuse this swerve off the original topic...

Graham, a space for lubricant in a plain thrust bearing makes sense, even if it results in axial float, However, with rolling element bearings, this float can be eliminated with impunity. I'm surprised by Emco's instruction, given that the FB-2's screws are supported by a pair of deep-groove ball bearings. Perhaps the instruction was to prevent gorillas applying too much pre-load to bearings with limited thrust capabilities. Any ideas?

Andrew. Backlash might encourage drill bit 'grabbing' in awkward materials, but I don't suppose many machine quill feeds have any such adjustment, so it can't be much of a problem in practice.

Thanks for explaining the tolerance spec. style; certainly odd...

Jason. The mental agility I meant was that required to read the dial, swapping between sixteenths and decimal inches (or mm!).

Edited By Kiwi Bloke on 02/07/2023 11:25:51

General point, emphasising what the wise Graham Meek wrote, a few posts back.

Feedscrew bearings, ideally, need to have zero axial float, should rotate smoothly, with no torque variation around the circle, and they must not bind (stick-slip phenomenon). It's all very well trying to get a low-friction bearing, but definite drag is needed, to keep the screw from rotating, as a result of vibration, and not helped by unbalanced 'balanced' handwheels. This 'unwinding' is a particular problem with coarse-pitch, small diameter screws (eg Myford cross-slide), because of the unfavourable helix angle. But it can also afflict screws with less steep helix angles, and provoked Graham M to make a lock for his Maximat's cross slide screw. It can therefore be useful to have a bearing arrangement that can be pinched up or pre-loaded to provide (consistent) drag. Apparently simple things can get complicated, when there are conflicting requirements...

A few loose ends tied...

I had hoped that the drawings would provide complete information for the pinion's manufacture, but that's all there is. As has been said, it looks like the teeth would have been shaped or planed, not cut with a circular cutter. It seems probable that the drawings were for Cowell's own use, and it was not intended that customers machine the pinion. I don't know whether any machining was required by the customer.

The drawing for the rack also omits the pressure angle, but does give an over-wire dimension. The shaft dia., into which the rack teeth are cut is 0.875" - 0.001 -0.002 (not sure I understand that tolerance, but that's what's written...); wire dia. = 0.083"; Overall dia. = 0.914" +/- 0.001". The rack teeth are indeed cut into the cylindrical surface, not a flat. I suppose this requires one less machining operation, but it seems a rather mean compromise. There was no provision for the pinion-rack backlash to be adjustable, the pinion shaft running directly in the main body casting, and there's not enough meat in the casting for an eccentric bush.

The dial was resettable, so requiring a known relationship to the pinion teeth was mercifully avoided. 24 pinion teeth, 40 rack teeth, so >1 rotation of the pinion possible. 48 divisions on the dial, so graduated in 1/16". Mental agility required...

Sorry, late to the party - had to do some serious excavating to find this. The following dimensional info is taken from an old Cowell blueprint, with my notes, in addition. Imperial dimensions.

The pinion, its shaft and the capstan boss are all machined from one piece of mild steel. Pinion OD 1.034 - 1.031". 24T X 1/8" CP. Wire = 0.083" dia. Measurement over wires = 1.099 - 1.097"

Hope this is enough info. PM me if you'd like a pic of the plan, but beware, the quality is rather poor.

Edited By Kiwi Bloke on 01/07/2023 07:08:22

Andy - found your blog and downloaded most of it for future reading and enjoyment. At first glance, all one can say is blimey! It's truly awesome work (in the true sense of the word).

It's just the sort of thing I'd like to see in model Engineers' Workshop. I think I will probably let my long-time subscription lapse - the mag has become so dumbed-down, and the wholesale lack of proof-reading annoys. There's amazing stuff to be found on the 'net - if one has the time to search - but there's little to amaze in MEW these days. Pity.

I think there's a risk of over-thinking this, and inventing new ways to polish a t*rd. Given the limitations of the rest of the machine, I'd suggest that the addition of a thrust washer either side of the bush is probably enough. Incidentally, I'd think the bush is cast iron or bronze (can't remember), not steel.

However, if you want to go all fancy, ball thrust bearings are available in 5mm ID, 11mm OD, 4 or 4.5mm axial thickness. IIRC, the bush is 12mm OD, so the bearing assembly can still be loaded into the machine from the right side. You'd need to shorten the leadscrew thread, and extend the end of the leadscrew, but you're committed to most of that already. You's also need a simple shield around each thrust bearing. I'd guess this would be the trickiest bit, but a stepped-bore tubular shield could be pressed or Loctited onto the non-moving part of each thrust washer.

Oh, forgot, Oriental bits ordered. No doubt the same approach would be possible...

Andy - interesting article, duly downloaded. Thanks for sharing it. I'll read it tomorrow, it's bed-time here! If we're branching out into casting entire headstocks - or more - there's also granite-filled epoxy (and other fillers) and concrete. Too many possibilities, not enough time...

Edited By Kiwi Bloke on 28/06/2023 12:27:14

May have misunderstood, but the 'proper' thrust bearing should be on the left side of the supporting bush.

Interesting that there's may be other possibilities for getting some new bits that can be modified for Emco machines.

Mentioning AKKO or Jupiter risks turning this thread into bayan porn... BTW, have you seen Ihor Pavlyk's YouTube videos?

Re reed manufacture. Take a look at **LINK**

John Cook, who seems to be the 'Great British Harmonica company', is a maker of harmonicas and works on other free reed instruments. His videos show the making of reeds from scratch, and the making of the punch tooling for it all. Impressive! Applicable to accordions...

Julius - Just read your reply. Understood. The wretched thing is making life difficult for you...

Just a thought - if you could improvise a temporary saddle stop, you could still turn the end of the leadscrew. Just slowly feed the saddle along the bed by carefully and sensitively pushing it by hand. The leadscrew will fit in the spindle bore.

Regarding your idea of having a larger diameter thrust face 'where the inner surface of a new, different handwheel will land on', just be careful. The face of the handwheel must run without wobble, but this is difficult to achieve if it's located by a non-precision, sloppy thread. The larger diameter, the more this is a problem. You may end up with a handwheel whose resistance to turning varies with rotation. That always feels horrible.

I would direct more attention to the right-hand end of the leadscrew and its bearing on the bush. I think Emco tended to make feedscrew thrust bearings too small in diameter. An interposed hard steel shim washer can work wonders. This is the thrust bearing that takes the cutting forces - the leadscrew is trying to escape to the right.

Julius - Just read your reply. Understood. The wretched thing is making life difficult for you...

Just a thought - if you could improvise a temporary saddle stop, you could still turn the end of the leadscrew. Just slowly feed the saddle along the bed by carefully and sensitively pushing it by hand. The leadscrew will fit in the spindle bore.