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.