The AJS V-Twin Crankshaft: Part 5 - Mainshafts

The nominal dimensions for the mainshaft lengths can be readily calculated from the locations required for the timing-side pinion to the camshaft drive on the timing side and the final drive sprocket on the drive side. The inside diameters for the mainshaft and big-end in the flywheels had already been determined and reamed to 29mm and 1.000’’ respectively.

I had also decided on a 4⁰ taper angle for the final drive sprocket and to not key the sprocket to the shaft. I discussed the question of keyed versus non-keyed final drive sprockets at length with Max Nightingale at Alpha Bearings many years ago and he was of the opinion that non-keyed is a better solution. Why? Well, because the loss of frictional contact surface area by the presence of a keyway on the tapered shaft is not compensated by the shear strength of a key. Velocette clearly came to the same conclusion in the 1920s because Mk1 OHC Velos also do not have a keyway to fix the externally-splined shock-absorber body to the crankshaft.

There are two main decisions that now need to be made before the mainshafts can be designed and manufactured. In no particular order, these are:

1)    Should the mainshafts be surface hardened?

2)    What should be the interference fit between the mainshafts and the flywheels?

I had originally planned to finish the mainshafts to size and then have them nitrided rather than going the more traditional route of using a carbon steel and then case-hardening the surface and grinding to size afterwards; this is why I had chosen EN40B steel. Nitriding is carried out at a much lower temperature of ~ 540 ⁰C rather than the ~800 ⁰C required for hardening carbon steel.

However, there was a flaw in this plan that I was not initially aware of. It turns out that, in addition to the desired diffusion layer of hard nitrides, the nitriding process also produces an unwanted “white layer” or “compound layer”. This white layer is undesirable because it can flake off and, because it is hard, could cause damage elsewhere in the engine and would also result in dimensional changes of the shaft. For anyone that is interested, there is more information here and here.

I contacted a supplier of gas nitriding and they stated that the thickness of the white layer with their process is typically 15 microns. This is 0.0006” in Imperial units and, on a shaft, would equate to a change in diameter of 0.0012”. As I was planning to have a tolerance of 0.0005” for interference fits and even less for the bearing fit on the shaft – a tight sliding fit - I decided not to use nitriding as a surface hardening process.

It is not critical that the crankshaft mainshafts are hardened. Unlike, say, gearbox mainshafts, they are not rotating with respect to another component and so in-service wear is not really an issue although if there is repeated assembly and disassembly then the shaft could wear. It is unlikely that the mainshaft and bearings would be repeatedly assembled and disassembled, however, there is also the final drive engine sprocket on a taper to be considered and for which hardening would be desirable.  

I had, in the meantime, investigated electroless nickel plating as a possible technique for increasing the surface hardness. This is an entirely different technique to electro-deposition of nickel and produces a thin layer of hard nickel by chemical deposition. More details can be found here. The hardness of the deposited layer depends on the composition of the aqueous solution used and, in particular, the phosphorous content - the less phosphorus the higher the hardness. Unlike electroplating, the layer formed from electroless plating does not flake off and has been used in many applications that require wear and impact resistance. Another major advantage of electroless nickel plating is that the deposited layer has a uniform thickness. Finally, heat treatment to increase the hardness of the deposited nickel is also possible.

Rather than send the shaft away to have it plated there are kits on the market that provide the chemicals and instructions to do this in the home workshop. I bought a kit from Delway Technical Services and started some experimentation. In terms of hardness, the deposited layer using this kit is quoted as 600 VPN200. This Vickers harness value translates to a Rockwell value of around 54 HRC. For comparison, the EN40B steel I am using for the shaft is delivered in the “T” condition which has a maximum value of around 32 HRC; my heat-treated O1 tool steel gears for the OHC drive are around 60 HRC. And so electroless nickel plating would increase the surface harness substantially and I would be quite satisfied with a value of the 54 HRC.

There are some additional bits of paraphernalia required in order to start plating – a pan to hold the parts (ideally stainless steel), a couple of litres of purified water, a heater to heat the liquid up to its optimum plating temperature of 90 ⁰C and a thermometer to check the temperature of the plating bath – see picture below.

I first made up a test piece of a similar sized round bar of EN40B to the mainshaft, supported on a mild steel rod, and put that in the plating bath for 1 hour.

The quoted deposition rate is 15 – 25 microns/hour and, as best as I could determine (I can’t measure to the accuracy of microns!) the diameter increased by 0.0015”. This equates to a surface deposition rate of 19 microns/hour. This is useful information because the immersion time can be used to determine the thickness of the layer that is deposited. I don’t have hardness testing at home and have to resort to trying a fine file on the surface for my very approximate “harness tester” but I could tell that the surface was substantially harder than the “as delivered” metal. On this basis, I decided to go the route of using electroless nickel plating for the drive-side mainshaft bearing contact surfaces and taper.

The second main consideration was the interference fit between the mainshafts and the big-end into the flywheels. I spent some time researching this and eventually came up with the nominal values of 0.0035” and 0.0025” for the mainshafts and big-end respectively. How did I arrive at these values? I talked with a couple of people whose opinion I respect – and they each came up with quite different values (!!) and by making calculations of the insertion force that would be required.

For those with a mathematical inclination, there is an analytical solution to the pressure between the contacting shaft and flywheel surfaces. Multiplied by the surface area of contact, this gives the force required to press the shaft into the hole. I have reproduced 4 pages of hand written notes on my calculations for both the mainshaft and big-end.

A couple of points to note: the coefficient of friction is an important parameter and much time was spent in searching available online literature to come up with an appropriate value; the necessary mechanical properties of EN24T (the flywheels) and EN40B (the mainshafts) are very similar and a simplification is then possible in the equations.

The required force to press home the mainshafts and the big-end came out to be around 12 tons and 7 tons respectively. My hydraulic press has a capacity of 20 tons and has a pressure gauge and so it will be interesting to see what pressure is actually required.

Drawings were made for the mainshafts, shown below

and machining was started. The picture below shows both shafts after completing the lathe work. There is excess material on the left side for holding and excess thread on the timing-side mainshaft to allow for a centre.

The surfaces have been machined to 0.008” oversize to allow for grinding. I have used an outside machine shop, Jayess Tools, for grinding as I do not have a cylindrical grinder in my workshop. Prior to grinding, oil transfer holes (one axial and 2 radial) for the oil feed to the big-end were drilled in the drive-side mainshaft and  tapped at the end to allow a high-tensile cap screw to be inserted for blanking off the drilling. An allowance of 0.002'' on the diameter (0.001'' thickness on the surface) was allowed for subsequent plating of the mainshaft.

The final operation was to use the electroless plating bath to put a hard nickel surface on the drive-side mainshaft. I did not want to plate the area in contact with the flywheel for the interference fit as this had already been finished to size and I found that workshop blue-roll (duct tape) can be used to effectively mask an area that doesn’t need plating.

This was then put into the plating bath for just over 1 hour

to deposited the required thickness of nickel

The mainshafts are now finished and ready for assembly into the flywheels.

 

The AJS V-Twin Crankshaft: Part 4 – Machining the Flywheels

Much of the AJS V-Twin project has been carried out during Covid lockdowns. When the first lockdown came in the UK in March 2020 panic erupted as the supermarket shelves were stripped of toilet rolls and baked beans. I also suffered a twinge of panic and immediately went to the workshop and made an audit of all the materials I thought I could possibly need but hadn’t yet ordered for the project – steel, aluminium, brass, phosphor-bronze, bearings, valves, tooling etc. As an afterthought I ordered 2 boxes of tinned haggis on ebay ....just in case; I still have 11 of the original 12 tins in my garage.

Part of this early order for materials included 2x 1900mm diameter x 40mm thick slices of EN24T for the flywheels and a 40mm diameter x 500mm long bar of EN40B for the mainshafts.

EN24T is a tough steel (Tensile Strength 850/1000 N/mm² and Yield 680 N/mm²) and has been used widely for crankshafts on performance engines. EN40B has similar mechanical properties and can be nitrided to give hardness of around 61 – 65 HRC. It was the intention to harden the shafts by nitriding but, for reasons that will be discussed in a later blog, an alternative route was followed.

The order of machining has to be carefully thought out to ensure that any particular machining operation does not make it difficult or even impossible for any subsequent machining. Before starting, I decided on the following processes for critical operations:

1)    I would ream the mainshaft and big-end internal diameters to get the accuracy and surface finish required. I do not have access to sufficiently accurate internal grinding - I don’t believe that my tool post grinder is adequate.

 2)    Accurate positioning of various holes would be set up using the DRO on the milling machine and then transferred to the lathe for main boring/reaming operations and setup with a dial indicator.

 3)    To ensure that the flywheels were identically machined they would be held together with accurate dowels through the mainshaft and big-end and then screwed together so that both could be machined simultaneously.

The first task was to face-off each flywheel in the lathe and to bore and ream the hole for the mainshaft.

The hole for the mainshaft is bored to 29mm minus 0.008’’ for reaming allowance and there is a 1/8’’ deep counterbore for a shoulder on the mainshaft. The first part of these reamers is tapered to facilitate easy entry and it is important that the reamer is inserted sufficiently that the entire hole is reamed with the parallel part of the reamer.

After reaming both flywheels, a mandrel plus nut and thick washer was made to hold the flywheel on centre so that the outside diameter could be machined. The nut and washer for the mandrel were made first so that the mandrel could be left in the chuck on-centre before machining the flywheels.

After machining the OD and facing off the other side on both flywheels they were put in the milling machine and 3x 5/16’’ BSF tapped holes and a pilot hole for the big-end were put into one flywheel and 3x clearance holes and a big-end pilot hole in the other.

A close fitting 29mm diameter dowel was also made for the mainshaft holes with which the 2 flywheels could be fixed together exactly on centre for further tolerance-critical machining operations.

The flywheel with the 3x threads was then carefully set up in the 4-jaw chuck on the lathe so that the face was perfectly orthogonal to the axis of the lathe bed and was centred on the big-end pilot hole.

The other flywheel was then centred and bolted to the first flywheel so that both big-end holes could be in-line bored and reamed.

There is a lot of off-centre mass rotating in the chuck for this machining operation and it was necessary to run the lathe at very low speed.

After boring the hole for the big-end both flywheels were reamed to exactly 1” ID

and after reaming both big-end holes, the clearance holes for the big-end nut/socket were bored

and the big-end boss was then machined on each flywheel

Each flywheel was then mounted on the rotary indexer on the milling machine and the balance holes were set up and bored. These were left undersize at this stage because the next step was to set up the flywheels on the faceplate so that each balance hole could be set up on-centre, in turn, and both flywheels bored simultaneously. The flywheels are doweled together through both the big-end and the mainshaft holes to ensure they are bored identically.

Each flywheel was then mounted on the dividing head on the milling machine to machine the annular segment.

Both flywheels were then again doweled together through both the big-end and mainshaft holes and the single hole opposite the crankpin was bored to 12mm. This will be used later for alignment of both flywheels for assembly in the press.

The last remaining major machining operation was to face-off the outside of each flywheel to form the mainshaft boss that will be in contact with the main bearings. Any minor adjustments for crankshaft end-float will be made immediately before assembly when all the parts have been made.

Finally, the oil drilling connecting the feed from the mainshaft to the big-end is required. This is a 2.5mm diameter hole and, rather than drilling this in the conventional way with a long drill, I chose to have it spark eroded rather than risk the drill wandering off and emerging in the wrong place.

The flywheel machining is now completed.

It is a somewhat academic point but the introduction of the oil hole creates an extremely small imbalance between the flywheels - one flywheel has the drilling and the other doesn't. This imbalance is referred to as a "Rocking Couple" and is an out-of-balance force along the axis of the crankshaft ....not to be confused with our behaviour back in the days of our youth...

The machining of the flywheels took me a few weeks and generated a few buckets of swarf and, because EN24 is tough steel, the machining ate through plenty of carbide and HSS cutters. The old adage “measure twice cut once” has never been more appropriate; I would not have wanted to have screwed-up the last machining operation!

I was pleased with the final result but the acid test will be when the crankshaft is finally assembled.

The AJS V-Twin Crankshaft: Part 3 – Design of the Flywheels

Before embarking on the detailed design of the crankshaft it was important to consider the information that was already known and any design or manufacturing constraints. Only then could further design decisions be made.

Firstly, the information that was known:

1)    The distance between the drive-side and timing-side main bearings in the crankcases: 4.128''. This determines the total width across the bosses of the assembled crankshaft.

 2)    Diameter allowed for flywheels (from crankcase machining): 7''

 3)    Width across crankpin of Harley Davidson big-end bearing: 1.76''

 4)    The Stroke: 86mm (apologies for mixed units!)

 5)    The inside diameter of the main bearings/mainshaft diameter: 1.125''

 6)    The masses of the various components (see previous blog) for the balance calculations

….and that was about it! Everything else is, as yet, to be determined.

90 years ago, manufacturers built jigs and fixtures for making components such as this and, once set up, this would provide repeatability and accuracy in spite of the relatively low intrinsic accuracy of the machine tools of the day. Today, a one-off crankshaft such as this and manufactured in a professional machine shop would be machined on a jig borer that, typically, would have accuracy of ~ 0.0001'' – 0.0002''.

I have a good but quite old lathe – a Harrison L5A that was manufactured in 1948 that I bought on ebay some years ago (I have the original receipt of purchase for £379 – 3 years before I was born! …this is less than I paid for it on ebay) and a Tom Senior Major milling machine of similar age but fitted with a modern 2-axis DRO. I also have access to a spark eroder with 50+ years’ experience and an excellent grinder both within a 20 minute drive away; although I prefer to do as much as possible myself, both were needed for some critical machining operations.

The accuracy required of the crankshaft is at the limit of accuracy that my ancient machine tools can deliver, which is about 0.001'' ….maybe 0.0005'' on a good day. The crankcases are very substantial (ie stiff) and the engine has 2 drive-side main bearings; this requires the crankshaft to have a runout of ~ 0.002'' maximum and there is a danger that tolerance build-up would swamp this. I will say more about how critical machining operations were dealt with in following installments.

I had decided to base the design, particularly for the balance, on using large diameter holes bored through the flywheels on the crankpin side. This had been used by Alpha Bearings for the AJcette crankshaft, below

And is also used on V-Twins; one example of a Vincent is shown here

Picture courtesy of Bonhams

Balance Factor

Pretty well all of the information on balancing motorcycle crankshafts is based on checking and correcting an existing crankshaft rather than designing a crankshaft from first principles.

Let’s start with the definition of the balance factor and how it is used. The objective is to balance the entire rotating mass plus a fraction of the reciprocating mass where the latter has been determined by a combination of analysis and experience that leads to an overall satisfactory level of vibration.

m_bal = m_rot + BF x m_recip

Where

m_bal is the total mass to be balanced

m_rot is the rotating mass

m_recip is the reciprocating mass

and

BF is the balance factor

Where the rotating and reciprocating masses are illustrated below

For a V-Twin engine, the rotational mass includes the masses of the big-end of both connecting rods plus the bearing and crankpin journal whilst the reciprocating mass includes the masses of the small ends of both rods plus both pistons (complete with rings, gudgeon pin and circlips).

Before proceeding too much further I had to decide what balance factor to use. The design of the frame and installation of the engine in the frame, for example, whether or not the engine uses head-steadies or isolating engine mounting affects the vibration response of the motorcycle and, in the absence of contemporary analytical tools that would be used by todays major motorcycle design departments, I have no other option than to pick a value and see how well it works. Modern production engines use balancer shafts that rotate at twice engine speed to eliminate, or at least reduce, 2^nd order vibrations in a plane orthogonal to the direction of piston motion. That is not an option here and so I have to choose a value from historical V-Twins that hopefully will provide acceptable balance.

The range that I have found for similar (ie 50⁰ or, in the case of Harley Davidson, 45⁰ V-Twins) is:

Vincent: 46% here

Vincent: 46% - 60% here

JAP  ~ 43% estimated but calculated “optimum” of 53% here

Harley Davidson 60% here

Harley Davidson 50% - 60%, 60% recommended here

I decided on a value of 60% for this engine. Why 60%? Well, you have to pick a value to proceed and with a value of 60%, if vibration is found to be unacceptable and a lower value would seem more desirable, it would be much easier to remove material from the side of the flywheels opposite the crankpin at a later date to reduce the balance factor rather than trying to remove more material from the big-end side to increase the balance factor.

The calculation of crankshaft static balance is, strictly speaking, not really concerned with weight (which is a force equal to its mass experiencing gravitational acceleration) but with Moments which, for static balancing, is the product of its weight multiplied by a distance. However, as gravitational acceleration cancels out in the equations used for static balancing I have here used the term Moment, somewhat loosely, to refer to the product of mass multiplied by distance.

In schematic form, the Moment to be balanced is equivalent to a mass of m_bal being hung from a location diametrically opposite the crankpin:

Moment = m_bal x Stroke/2

which is equated to the moment of the balance holes to be machined in the flywheels, illustrated below.

Moment = m_bal x Stroke/2 = ∑ m_hole x h_hole

The next step is to calculate the diameters and locations of the balance holes. This required numerous design iterations of the locations and dimensions of holes. This calculation must be done with care; it is not a difficult calculation but it does require careful bookkeeping to make sure there is no double counting. For example, the mass of the crankpin that has been weighed already includes the mass that would be occupied by that part of the crankpin that is contained within the flywheels and this must be subtracted from the total crankpin mass. There is a boss at the crankpin and a clearance hole for the crankpin nut. And, of course, there are 2 sets of holes per flywheel and there are 2 flywheels…; this must all be accounted for properly.

I won’t go into the lengthy and somewhat boring details here except to add a few points:

1)  I decided to use a straight-fitting crankpin rather than the taper fit that came with the HD connecting rod and big-end. The original and reground big-end pins are shown below. Why? Well, I don’t believe that I am able to machine the female part of the taper within the flywheels to the level of accuracy required for good alignment. This hole, and also the hole for the mainshafts, will be reamed (note: the crankpin was ground after the flywheel holes had been reamed to ensure the desired interference fit).

 

2)  After making detailed drawings I decided that the diameter of the hole required to fit a socket over the existing hexagon big-end nuts would be too great and would reduce the wall thickness between the big-end nut clearance hole and the mainshaft to be unacceptably small. To reduce the diameter of this hole for the big-end nuts the nuts were modified from hexagons to dodecagons (12 sides), which can be seen in the above picture, and a copper electrode was made to spark erode 2 dodecagon sockets for subsequent assembly. The electrode (after use) and the sockets are shown below.

The ratio of a circumscribed circle of a hexagon to that of a dodecagon is 1.116. With the dimensions here, this allows a hole that is 0.18'' smaller diameter by using a dodecagon and, in turn, that increases the wall thickness between the mainshaft hole and the clearance hole for the nuts by 0.09''.

 3)  It was not possible to satisfy the amount of mass to be removed by simply boring holes. On a single cylinder engine this would not have been a problem but as 2 connecting rods and 2 pistons are included in the balance calculations for a V-Twin, additional mass needed to be removed and at a significant distance from the crankshaft centre to increase the moment. To this end, an annular segment was introduced at the flywheel periphery.

The schematic and table below show the dimensions of the holes and their relative contribution to the overall balance.

 

	Diameter (mm)	Contribution to Balance %
Hole #1	36	8.5
Hole #2	36	51.1
Hole #3	12	8.5
Segment #4		37
Hole #5	12	-5.1

 

The additional hole, #5, has been introduced for manufacturing and assembly purposes as will be described later.

The overall balance factor with this setup worked out at 59.2%. It will be interesting to see how it turns out in practice after everything is made and assembled.