The DOHC 250 Velo Engine: Inside the Cambox

I have intentionally not disturbed the cambox up to this point in the engine build. Why? Well, because it appears to work just fine and there has been so much else to do but, as now the engine build is nearly complete, it was time to look inside.

This is what it looked like before stripping.

However, before dismantling the cambox I figured it would be a good idea to check the “as-received” valve timing. The relative valve timing, ie the relationship between the Inlet and Exhaust is determined by the position of the gears on the camshafts at the ends relative to each other and I have assumed that the setup reflects how it was run on the engine.

The absolute valve timing (ie the actual valve timing) is governed by the positioning of the bevel gears. Now, the large top bevel gear that is on the end of the shaft driving the other gears – part number K-18, has 44 teeth. One complete rotation of this shaft equates to 2 revolutions of the crankshaft and so the crank-angle change that would result from changing the engagement position of this gear by one tooth would be (360⁰ x 2) / 44 which equates to ~16⁰ crank angle. This would be a very significant change and this helps in determining the absolute timing by avoiding unfeasible extremes.

After a couple of changes to the relative positions of the bevel gears I came up with the following valve timing:

IVO 60⁰ BTDC

IVC 70⁰ ABDC

EVO 70⁰ BBDC

EVC 30⁰ ATDC

The 2 more reliable figures are EVO and IVC as the other readings can be affected by the loading of the other cam during valve overlap.

These can be compared with valve timings on other engines

and also indicates that changing the bevel gear by one tooth would put the valve timing outside the range used on any of the engines listed above.

After removing the outer steel plate supporting the intermediate gears all of the gears can be seen more clearly.

I have marked the tooth engagement of all gears (and also the bevel gear engagement) so that they go back in exactly the same place. The position of the intermediate gears is not actually critical as they have been set up with hunting teeth. The number of teeth on the gears from the camshaft on one side to the camshaft on the other side is:

30 – 31 – 32 – 30 – 32 – 31 – 30   

After removing the gears (which are all bushed with an oilway) the main structural support plate is revealed.

I initially wondered where there were 2x 2BA threaded holes in the plate that seemed to have no obvious purpose but it soon became clear; they were there to be able to insert 2 screws to lift out the steel plate from the aluminium casting.

After removing 4x ¼” BSW Allen screws and heating the entire structure, the plate plus everything attached to it was separated from the outer casing.

Both cams had minimal wear

but I replaced the balls (10x 7/32” diameter) in 2 of the Hoffmann bearings

which were showing a bit of play. I measured the base circle diameter at 1.008” and the lift + base circle came to 1.315” which gives a valve lift of ~ 5/16”.

All of the parts were in excellent condition, however, there was one conundrum, namely: what was the lubrication strategy? There are clearly 3 different ways that oil can flow in or out of the cambox. The oil drain is obvious – the pipe at one end of the cover, shown in the picture below.

An oil feed to the cams is also obvious, indicated by the arrows in the picture below.

There are quills of ~2mm diameter at the ends of pipes that would supply oil directly to the cams. However, it is questionable how much oil would be delivered using this method; the oil would have been provided from the upper bevel chamber via a slot in a K-12 phosphor-bronze bush in the same way as a standard “K” cambox oil feed and although there is an oil seal in the centre and immediately behind a bearing (to prevent the oil simply discharging into the cambox at the point of entry). I am sceptical that this approach would provide the amount of oil required to lubricate the cams effectively and, indeed, the entire gear train. This also indicates that the cambox was originally designed to work with an internally pressurized system rather than a targeted distribution approach.

I suspect that whoever designed the cambox must either have had the same concerns or, alternatively, the cambox started life on a pressurized “K” type system but transitioned to a strategic lubrication system (as on KTT 581) because that would explain why there are oil feeds that are positioned at the ends of the cambox. Alternatively, these could have been additional oil drains to lower the level in the cambox when the engine was stationary to prevent it leaking away via the pushers and covering the cylinder head in oil. The inside and outside (it is a 1/8” BSP thread) of one of these is shown in the picture below.

With the lubrication system that I have decided to use, the central feed system is redundant – the upper bevel chamber will not be pressurised and, instead, I will introduce a quill into both ends which will be fed under pressure directly from the oil pump.

Similarly, the upper bevel will also be provided with a direct oil feed ….more next time.

The DOHC 250 Velo Engine: The Final Build – Part 1

Before starting the final engine build there were a couple of jobs that I had postponed, simply because up to now there was always something more important to do.

The first of these was to replace the small end. The small end bush that had come with the MOV flywheels/connecting rod assembly had just a bit too much wear and so a new phosphor-bronze bush was made on the lathe and milling machine,

an oilway along the length of the bush put in with a long 1/8” end mill

and then pressed in and reamed to fit the gudgeon pin that had come with the engine.

The second outstanding task was to check the crankshaft balance. I would assume that my Ebay-acquired MOV crankshaft would have been balanced to an MOV piston and the only difference is that I am now using the steeply domed race piston that came with the DOHC engine. Nevertheless, the balance needed to be checked. The crankshaft must balance all of the rotational mass + a fraction (the Balance Factor) of the reciprocating mass.

The reciprocating mass is the mass of the complete piston assembly, m_p, plus the small end, m_se.

By mounting the crankshaft assembly so that it can rotate freely and adding weights to the small end the balance mass, m_bal, can be determined.

First, the small end was weighed

and then the piston + rings + pin + circlips, m_p, (which came to 347g) before the flywheel assembly was mounted in a pair of used main bearings on vee blocks on the milling machine table.

I have used this arrangement rather than the crankshaft balancing jig as the assembly can be rotated without the end of the conrod hitting the deck!

The first step was to ensure that the existing balancing puts the big end at the top, rather than one side, when rotated.

It does. And so, the next step was to add weights in a small polythene bag to the small end so that the crankshaft would stop in any random position, without any preference, when rotated.

The mass added to do this, m_bal, was 214g. The balance factor, BF, is then defined by:

BF x (m_se + m_p) - m_se = m_bal

Which gives a balance factor of 73%

Is the acceptable or not? It is not easy to find documented “hands-on” experience of balance factors for MOVs but I did come across one article titled “Making a MOV Go” on page 6 of Fishtail 18 (October 1959) which stated:

 “ ….to enable a balance factor of 70% to be achieved. No experiments were made with other balance factors, the motor proved smooth between 4000 RPM and peak revs of 8500-8700 RPM.”

and so the value of 73% in the DOHC engine would seem to be just about right.

I have seen extremely high values of 85% quoted for MOV balance factors but I can find no evidence to back this up. If I substitute the raw piston weight of 8 ¼ oz (235g) for a 6.5:1 compression ratio piston that is given in the Velo technical data here then this would give a balance factor of 80% and this value would be reduced if the slightly heavier 10:1 race piston data were used. For the record, the raw piston weight of the DOHC piston is 307g.

I had already sorted out the lower bevel gear meshing (see here) and now the solid steel dummy “bearings” that I had been using were removed, the oil pump was inserted into its hole, the new (modified) main bearings were fitted, the crankshaft was put in and the crankcases sealed – exactly the same procedure as described earlier for KTT 305. Prior to assembly, the modified Woodruff key that engages with the lower bevel gear was heat treated.

The K-45/2 timing gear cover plus all the internal bits and pieces were assembled. 2x 1.25mm thick steel discs were placed behind the oil pump driving piece to get good engagement of the K-34 and K-72 gears as described in a previous blog. There are 2 different gaskets (K-68 and K-68/2) depending upon the timing crankcase/timing gear cover combination – needless to say I had the wrong gaskets in stock and so had to make the other one.

I have also fitted a new spring and ball to the oil pressure adjuster.

With the changes that I have made to the lubrication system these will no longer be used to adjust the oil pressure but they do need to be there to block the hole that connects the chamber to the feed side of the oil pump.

The inner timing case then needed modifying

to enable it to fit over the boss that I had attached for the crankshaft quill oil feed.

If the magneto chain ran in an oil filled chamber then there would be a problem ….but it doesn’t.

I had checked the piston, piston ring and bore dimensions some time ago and found clearance of 0.008” on the thrust/antithrust faces at the bottom and the 2 compression rings had a 0.032” gap. Recommended values for these quantities are generally quoted “per inch of bore” and to be able to make comparisons my measurements equate to 0.003” and 0.012” per inch of bore respectively.

So, what values are recommended? That depends where you look…. For engines in standard trim (ie, non-racing) Hepolite/Heplex recommended a comparable piston/cylinder clearance of 0.0015” per inch of bore (see eg The Vintage Motorcyclists’ Workshop, Radco, page 53). On the other hand, Phil Irving (page 68 of Tuning for Speed) gives a value of 0.0027” – nearly double. The piston/bore clearance dimension on this engine is very close to Phil Irvings value (0.003" versus 0.0027”)

For the compression ring gap, Radco quotes a value of 0.012” (which, for comparison, equates to 0.0045” per inch of bore) for a Velocette KTT whereas Phil Irving suggests 0.010” per inch of bore – again, double.

I checked the ring thickness against published Velocette MOV data and they were the same and so I bought a set of new rings. These fit perfectly and gave an “off-the-shelf” value of 0.025” for the compression ring gap. This equates to 0.0093” per inch of bore which is pretty well exactly Phil Irvings recommendation.

So, it would seem that the clearances in this engine conform pretty closely to those recommended by Phil Irving.

Incidentally, I was amazed that I could simply phone Cox and Turner

and buy a set of NOS (New Old Stock) piston rings that arrived the next day for a fairly obscure British motorcycle that ceased production in 1948!

The next step was to put the piston + rings on the connecting rod and put on the barrel and cylinder head. The circlip grooves in this piston are flat-bottomed and so new Seeger circlips were used to hold in the gudgeon pin. There are not that many occasions when I need circlips but I bought a box of assorted imperial-sized circlips many years ago

and these do come in useful when needed.

The barrel and head had been given a coat of cylinder black before assembly. It was now time to set up the vertical drive.

The drive shaft (which would usually be part number K-49 on an early “K” engine) that came with engine is substantially longer, 7/8” longer in fact,

and although I don’t have a Mk 2 shaft in my workshop (except in the engine of the bike I recently purchased) my guess is that this is a shaft for the later engine. It also looks to be new with only minor witness marks of having been used.

This required a pair of Oldham couplings - the question is how thick should the flange on the couplings be? The shaft was put in place and a bunch of washers + shims + a feeler gauge placed between the shaft and the drive at the lower end

to find out. 2 new Oldham couplings were then made with a slightly thicker flange (0.218” here vs 0.188” for the thickest commercially available part (see here)) from O1 tool steel,

finished to size and heat treated

before cleaning up

and fitting. In the above picture, the 2 new ones are on the left and, for comparison, a standard “K” coupling on the right where the difference in flange thickness is apparent.

I have used a slightly higher tempering temperature than usual (260⁰C) for the heat treatment to trade off some hardness for additional toughness. The thicker flange should also help in that respect.

A check after assembly showed 0.008” vertical clearance

which should be fine.

Much of the engine build is now complete. The main outstanding tasks are: check the cambox internals, complete the lubrication system and fit the magneto.

And then it will be Christmas and I will be banned from the workshop for a few days.

The DOHC 250 Velo Engine: Securing the Cylinder Head and Cambox

Having completed the eccentric studs, the next step was to check the compression ratio – up to now I have had no means of bolting down the barrel and head properly.

The crankshaft assembly and piston + rings (the old rings – I have a new set but I’ll come on to that at a later date) were reassembled into the crankcases. The valves, without springs, were then inserted into the head and the head bolted down and checked for clearance at TDC. Plenty of clearance here - there would be no chance of the valves and piston making contact during the valve overlap period so the head was reassembled,

the piston was positioned at TDC

and with the engine held on its side with the spark plug hole uppermost, the combustion chamber space was filled with oil (R 40) from a burette.

The combustion chamber volume was measured to be 30.5cc which gives a compression ratio of slightly over 9:1. My guess is that this is probably around the value of when the engine was originally built and, although fairly high, should be OK with modern 98 RON pump fuel.

The next step was to determine the dimensions of the cylinder head studs, which also provide pillars for the cambox, and the cambox securing bolts. Like the eccentric studs for the cylinder, these were made in EN24 steel for strength and took quite a few hours of machining, for example, the 2 long bolts on the drive side of the cambox

and milling multiple hexagons.

Anyway, they all worked out well and the cylinder head and cambox fitted perfectly.

At this stage, I had intentionally left excess material on the studs/pillars so that there was a gap between the valve pushers in the cambox and the valves.

The height of the studs was then reduced to give a running clearance between the pushers and the valve stems with the cambox seating properly on its 4 contact points. I have set the clearances to 0.020” on the exhaust and 0.012” on the inlet. With care, I am hoping to build this part of the engine without resorting to shimming.

The final step was to chemically black all the bolts and studs ready for final assembly.