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Technical Information (MGB 24)


Added - 2/19/01

View of Assembled components (click for enlargement)

Complete Disassembled Kit (click for enlargement)

Assembled Push Rids with Upper Push Rod Spring Retainer Protruding Above Engine Block (click for enlargement)

Assembled Push Rods with Lower Push Rod Spring Retainers (click for enlargement)

Some people may say that this system is just nothing more than a "Rev Kit". However, as you read on, you will see that a lot of thought, not to mention a lot of time and money, went into the finished product..

First some background information on why we embarked on this project in the first place.

Back in 1993 I read an article in the October edition of "CIRCLE TRACK" titled "Trulink Rocker Arm System". This article dealt with a company called DECUIR DEVELOPMENT and their development and adaptation of a new valve train system.

In April 1994 "CAR CRAFT" also published a further article featuring this system titled "Spring Ahead".

Back then "PETERSEN PUBLISHING CO" which is now called "emap -petersen" published both of these magazines. For further insight on this valve train system, request copies of these articles by sending check or money order for $5, made payable to EMAP. Address to: "Emap. USA. Inc", Attention Editorial Assistant, 6420 Wilshire Blvd., Los Angeles, CA 90048-5515. Phone 323 782 2000.

What follows are the reasons for, and a brief description, of this system.

During the engine rebuild process it will be noticed, after installation of the cylinder head and rocker gear, that turning the engine over by hand has become increasingly more difficult, all in part due to valve spring #s. Simply put, it requires horsepower to turn the valve train. The greater these valve spring #s, the greater the horsepower required to turn the valvetrain.

All assembled valve spring #s are subject to a rocker arm multiplication factor. This multiplication factor is exactly what the camshaft lobe sees at the lobe/lifter interface.

For our example let us use an MGB engine valve train with British Automotive's 1.55:1 roller rockers, camshaft P/N 6710-18) and Moss Special Tuning valve spring set P/N TMG10707 (423-455).

First, we need to observe the following data:

Valve spring 140# (outer)
Valve spring 76# (inner)
Valve Spring 216# (combined)

Valve spring seat 67# (outer @ installed height of 1.575")
Valve spring seat 37# (inner @ installed height of 1.575")
Valve spring seat 104# (combined)

Rocker arm ratio 1.55:1
Camshaft lift .289"
Valve clearance .016"

Theoretical valve lift .289" X 1.55 - .016"   =   .432"
Valve spring # (Full lift @ 216# X .432")   =   93#
  (Valve spring seat #)   =   104#
  Total     =   197#
Camshaft lobe load @ .432" valve lift (197# X 1.55)   =   305#
Total camshaft lobe load per revolution (305# X 8)   =   2440#

Now let us look at what happens if we move the inner valve spring 37# to the pushrod side.

Valve spring seat 67# (outer @ installed height of 1.575")
Valve spring # (Full lift @ 140# X .432")   =   60#
  (Valve spring seat #)   =   67#
  Total     =   127#
Camshaft lobe load @ .432" valve lift (127# X 1.55   =   197#

We now have to calculate the pushrod inner spring #. Let us presume that we have the same valve spring seat 37# at the same installed height 1.575". Since we know that this spring rate is 76# we can simply multiply this number by the camshaft lift of .289 (76# X .289") = 22# then add the seat # for a total 59#.

Camshaft lobe load per revolution (59# X 8)   =   472#
Camshaft lobe load per revolution (197# X 8)   =   1576#
Total camshaft load per revolution   =   2048#

In one complete camshaft revolution we would have realized a savings of 392#s. (2440# minus 2048#). At 1000 engine RPM (500 camshaft revolutions) we would see a savings of 196,000#s.

NOTE: This is just an exaggerated example and not a working model. In this case the pushrod spring seat# would have been far too excessive to a point where the lifter, in all probability, would not rotate, resulting in rapid camshaft lobe and lifter wear.

The forces present at the cam/lifter interface are the sum of the inertia force and the valve spring force. At low engine speeds the inertia force is very small so almost all of the force on the cam is due to the valve spring. At maximum RPM, the valve spring force is just equal to the inertia force before the valves float. This means that friction savings would decrease as the engine speed increases.

Reduced valve spring #s is not the only advantage with this system. We continue with the following:

During the valve closing sequence, with the valve fast approaching contact with the valve seat, we have linear velocity and momentum of the valve, pushrod and lifter. Likewise, we have angular velocity and momentum of the rocker arm. When the valve contacts the valve seat, it's stored energy is taken up by the valve spring. The stored energy in the rocker arm, pushrod and lifter is however not taken up.

This energy is stored mechanically in these 3 components, especially in the pushrod. The faster the valve is moving, the instant before it makes contact with the valve seat, the more energy will be stored in the valve train. When this energy is released the rocker arm tends to "dance" on the valve stem tip. The more energy that is in the system, the more force the rocker will have to "tap" the valve with.

The "Decuir Development Company" found a way to absorb the stored energy of these 3 components. This was done by positively linking the rocker arm and pushrod assembly as a single unit (by way of a special rocker arm adjusting screw and a threaded female pushrod sleeve). Also by preloading the lifter with the installation of an inner valve spring on the pushrod itself.

The above text was edited from the original format with changes and additions made where necessary.

Let us continue with the trials and tribulations of attempting to adapt this system to the MGB engine (June 1995).

First, this system was not going to work unless we could find some way of straightening up the inherent "inward" pushrod lean common on all "B" series engines. Manufacturing the following components partially solved this problem. The maximum amount we could offset the rocker pedestals was found to be .110".

New 6061-T6-aluminum rocker pedestals (4) with maximum offset .110".
New cylinder head studs (4).
New 1.55:1 roller rockers (8). To allow for .110" offset.

Next up was to order the components for one valve only. This consisted of the following components:
Special pushrod.
Rocker arm screw-in stud.
Female pushrod threaded sleeve.

These components were assembled, without the pushrod spring, as per illustration "A".

Maximum valve lift readings were recorded and compared against previous engine rebuilds that included the same camshaft and the OEM rocker/pushrod set-up. Something was drastically amiss; we could not achieve anywhere near the valve lift of an OEM 1.426:1 * rocker arm. Remember we are using a 1.55:1 roller rocker arm.

*Most camshaft regrinders use the OEM ratio of 1.426:1 (1.358" divided by .952") to quote anticipated maximum valve lifts. These numbers represent the center line distances of the valve contact anvil and the valve clearance adjusting screw relative to the rocker arm bush center line. It is quite possible that you may come up with a ratio somewhat less than 1.426:1. Simply put, OEM rocker arms never give you this ratio. The 1.358" number has been verified over and over again, however, the .952" number can vary as much as +.030" between various rocker arms.

Since we were using a 1.55:1 modified roller rocker arm (1.443" divided by .931") we should have seen increases in maximum valve lift. Apparently, the critical point is in the initial pushrod inward lean, valve closed, and the pushrod to rocker arm attachment method. The engine was turned over slowly and the following observed, as the pushrod moved upward it began outward lateral movement to a point before full valve lift. No more lateral movement of the pushrod was observed until the same point was reached on the valve closing cycle, at which point the pushrod began inward lateral movement back to the original starting position.

Another pushrod was manufactured that would allow the use of the OEM adjusting screw, male and female threaded adapters as in illustration "B". Once again we went through the checking procedure mentioned above. On this particular check we ran into the same problems but to a lesser degree. Pushrod lateral movement was certainly less than "A" however, we were still short on valve lift.

At this stage we simply decided to fit an OEM style re-profiled rocker arm as per illustration "C" with the same pushrod as used in "A" without the threaded female adapter.

Hey Presto! Problem solved. Very little pushrod lateral movement with anticipated valve lift.


The OEM rocker arm with its ratio of 1.426:1 allowed the pushrod to attain a more vertical position at rest. I doubt whether or not this was the only reason, especially when look at the two dimensions, for each rocker, from the rocker arm adjusting screw ball center line to the center line of the rocker arm. 1.55:1 roller rocker arm .931" 1.426:1 OEM rocker arm .952" a difference of .021".

Having the pushrod and rocker arm assemblies independent of each other ("C") has more geometrical advantages than either "A" or "B".

During the valve opening and closing sequence the rocker arm is constantly changing angle. In "C" the rocker arm adjusting screw ball and pushrod cup work independently of each other, this allows the adjusting screw ball to establish any unrestricted angle during valve opening and closing without any great amount of pushrod lateral movement or loss of valve lift.

In "A" and "B" the adjusting screw and the pushrod are an assembly. Although there is full ball and cup articulation within this assembly, I am at a loss to explain why we could not achieve greater valve lifts. I must admit I don't quite understand the geometry at work with this system, but obviously it has something to do with pushrod installed angles and the accompanying rocker arm angles during the valve opening and valve closing cycles.

Decision time. Where do we go from here? We decided to continue with the project but leave the rocker adjusting screw and pushrod as independent components. We now concentrated on developing the rest of the system. Before valve spring and pushrod spring analysis could be carried out we needed further information.

All the valve train component weights were recorded. Our intended valve spring set P/N TMG10707 (423-455) was sent to "Elgin Racing Cams" to establish "natural frequencies" (vib/min) and "harmonic numbers" as was a roller rocker arm to establish the "moment of inertia" scale reading.

What came back was a full analyst of the valve train requirements to allow the engine to operate safely in the 6500-RPM (30% safety factor) to 7000-RPM (15% safety factor) range when using this new system.

Our original intention was to just simply install the 423-455 valve spring set and use the outer valve spring on the valve and the inner valve spring on the pushrod. Steve Gruenwald of "GRUENWALD SOFTWARE" in conjunction with Dema Elgin came up with some interesting conclusions. If we had split up the valve springs, as I intended, we would have had the valve springs vibrating out of control well before 6500-RPM with subsequent valve float. Apparently, the "natural frequency" and "harmonic number" range numbers at 6500-RPM of both springs was too low when coupled with the higher "moment of inertia" (more weight centered over the roller rocker tip than the OEM rocker) of the roller rocker assembly.

New valve springs were sourced resulting in the following:

Valve spring seat
(Rated @ 193#)
77# (Outer @ installed height of 1.552")
("Natural Frequency" 27,803 vib/min)
("Harmonic Range Number" 8-9th)
Pushrod spring seat
(Rated @ 20#)
20# (Inner @ installed height of 1.240")
("Natural Frequency" 32,804 vib/min)
("Harmonic Range Number" 12-13th)
  Total   97#
Valve spring - Valve open 158# (Outer)
Pushrod spring - Valve open 50# (Inner)
  Total   208#

As previously mentioned, this set up allows us to operate at an engine speed of 6500-RPM with a 30% safety factor. This will also allow enough spring pressure to account for increasing valve train flexibility as the engine speed increases without having excessive pressure that causes premature cam/lifter wear.

Let us look at the following:

Outer valve spring # (valve open) subject to rocker arm multiplication ratio of 1.55:1. (158# X 1.55)   =   245#
Pushrod spring inner # (valve open) not subject to rocker arm multiplication factor but subject to camshaft lift.
  (Pushrod spring 105# X .289")     =   30#
  (Pushrod spring seat #)     =   20#
  Total     =   295#
Total camshaft lobe load per revolution (295# X 8   =   2360#

When compared with our example of the OEM dual valve spring set 423-455 we realize a savings of (2440# minus 2360#) = 80# per crankshaft revolution. At 1000 engine RPM (500 camshaft RPM) we would see a savings of 40,000 #s. However, as we stated previously this savings would not be linear.

Let us presume that this figure of 40,000 #s is our maximum savings at 1000 engine RPM and using 6500 engine RPM as maximum engine RPM "valve float" condition with zero #s load on the camshaft lobes, can we just do a little math and draw a straight line graph representing theorized loads at various engine RPMs?.

1000 RPM = 40,000 #s/min savings    (Idle Speed)
2000 RPM = 33,300 #s/min savings
3000 RPM = 26,640 #s/min savings
4000 RPM = 19,980 #s/min savings
5000 RPM = 13,320 #s/min savings
6000 RPM = 6,660 #s/min savings
6500 RPM = Zero #s/min savings    (Valve float)

This should give us some approximation of spring # savings at various engine RPMs.

Another benefit of separating the inner and outer valve springs lies in the elimination of friction between these two springs. This friction would have increased with engine speed. Not having two springs rubbing against each other (or a damper) reduces the wear of the springs.

Yet another benefit is provided by the pushrod spring which permanently preloads the camshaft. This preloading dampens the camshaft cyclic loading and unloading forces, resulting in a more stable valve train with reduced distributor drive gear "chatter" and stable ignition timing.

Less valve train loads equals less valve train component wear.

With this set-up on my 1979 MGB "Limited Edition" the most noticeable thing, upon engine start up, was how much quieter the engine sounded.

On June 3rd 1996 we drove the MGB out of the shop (mileage 104,730. As of November 11th 2000 (mileage 110,429) with only 5,699 miles the AVT system is still working fine.

The engine in the MGB is my own developed 1924cc (actually 1927cc) with 9:1 GCR. This engine was previously dynoed at a whopping 137ft/lbs of torque @ 3500 RPM and 115BHP @ 5000 RPM. Although the AVT system was designed to peak at 6500 RPM with a 30% safety margin, I see no need to rev the engine above the maximum BHP RPM. Occasionally, when the situation deems necessary, I jump on it, with no complaints from the engine.

Did I notice any improvements in the cars performance after fitting the AVT system? I can honestly say I don't know. With such a flat torque curve 128ft/lbs @ 2500 RPM. - 127ft/lbs @ 4500 RPM prior to the AVT installation, significant increases would have to have been realized in both torque and BHP figures for a "seat of the pants feel".

Very few MGB owners actually understand that valve spring # selection should be matched to the intended engine operating RPM range. Camshaft dynamics require that low engine speeds be avoided when using heavier than stock valve springs. That is why camshaft regrinders require that you run the engine, upon initial start up, at 2000 to 2200 RPM for 20 to 30 minutes to aid with the break-in period of the camshaft lobes.

I have a set of valve springs used for racing which come with a warning that the engine should not be operated below 3000 RPM.


Prepare your engine block as you would do in a normal engine rebuild, however, have your machine shop bore down the pushrod guide tube holes (using a cutter the same size as the hole) to within 1" above the lifter bore. This will ensure that the pushrod spring will have sufficient wall clearance.

Also, the engine block deck height needs to be measured at the front and rear then recorded per the illustration below. OEM engine block height was approximately 9.941"

A. Cylinder block mean measurement + 1.136" = ????????

Where possible, but not essential, all the camshaft lobes should be of the same heel diameter within .005". Reground camshafts are notorious for having large variances in heel diameters. For your information the OEM camshaft 88G303 used from 65 thru 74 had a heel diameter is 1.120".

Our engine block deck height was recorded at 9.921" .020" below the OEM 9.941"
Our reground camshaft heel diameter was recorded at 1.030"
Some .090" smaller.

We used the above information to establish where the lower pushrod spring retainer groove should be machined on the pushrod to establish a seat pressure of 20# at an installed height of 1.240".

In our situation we machined the pushrod groove 2.800" from the base of the pushrod. Arriving at this measurement was very time consuming, however, it established a base from which to work from for future AVT adaptation to other engine blocks.

It is safe to assume that most MGB engine blocks are going to be redecked during the engine rebuilding process. The amount of material removed from the engine block directly effects the pushrod spring installed height. For example: if we removed .010" from the surface we will have decreased the pushrod spring installed height by the same amount, which then would increase the seat #. When using the 105# rated spring we would see an increase of approximately 1#, raising the seat # to 21#. So, for every .010" removed from the engine block will increase the pushrod seat # by 1#.

Reduced camshaft lobe heel diameters have the adverse effect. Any diameter smaller than the 1.120" OEM must be subtracted from the OEM diameter, then divided by 2. Our reground camshaft measured 1.030" a .090" difference, when divided by 2 we have .045" difference. Here in our example, if we had not compensated for this difference, we would have a reduced seat # of 4.5#, for a total of 15.5#. So, for every 020" reduction in lobe heel diameter will see a reduction in seat # of 1#.

We understand that many MGB engine rebuilders are not too concerned with the uniformity of matching camshaft lobe heel diameters. We decided it would be more convenient to standardize where the lower retainer groove is machined on the pushrod. All pushrods will now have this groove machined 2.800" from the base of the pushrod.

Changes in the spring seat # will now be controlled by the height of the upper pushrod spring retainer above the engine block deck. In our previous example, to attain a pushrod spring seat 20#, we needed the retainer to protrude above the deck by approximately .250".

What we have now done is to convert all the calculations back to the OEM engine block deck height (9.941") and the OEM camshaft lobe heel diameter of 1.120". We have also made allowances for the compressed cylinder head gasket of .023".

Using this information, we would require that the upper retainer protrude .214" above the engine block deck. Example: 214" minus .023" = .191" multiplied by spring # 105 = 20#. For an installed height of 1.240".

The upper pushrod spring retainers will be supplied in sufficient length for machining purposes. This would then become a simple matter of measuring each individual retainer, without the cylinder head gasket in position, at .214". Remember to mark the correct location (1 thru 8) of each retainer.

When assembling the lower pushrod spring retainer to the pushrod, we recommend that the retainer be seated on the 2 split collars by inverting the pushrod and, while the retainer is supported, tap the lifter end of the pushrod with a non-metallic hammer.

Increasing rocker arm ratios, whether they be OEM rocker arms or roller rockers, are established by reducing the center line distance between the valve adjusting screw ball and rocker arm bush. This only holds true when using OEM rocker arm pedestals.

The roller rocker arms that were used in the AVT system were offset a further .085" at the valve side of the rocker arm. The rocker arm pedestals were offset .110". This .025" difference had little effect on the rocker arm roller to valve stem centerlines. We would have liked have offset the roller rocker arm .110" but unfortunately, this .085" was the maximum offset the rocker arm design would allow.

Prior to completing the installation of our AVT system, we checked out various roller rocker arm ratios. Due to pushrod lean we found that the maximum ratio we could use was 1.59:1 (1.443" X .905"). Using 1.59:1 roller rocker arms over the 1.55:1 set would result in a maximum valve lift difference of approximately 2.5%. However, this 2.5% would be applied throughout the entire valve lift.

Depending on the CFM flow rates of the intake or exhaust valve, increasing the valve lift maybe worthwhile.

Having spent around $1200 to develop this system for the MGB 5 main bearing engine, I have to be realistic as to the marketing potential of the AVT system to the MGB owner. Since an investment in a 1.55:1 roller rocker kit will set you back $375, we need to determine what other costs will be incurred to purchase the rest of the kit.

With all the prototype production costs taken care of, we can now estimate what it will cost to put together the remaining components based upon a quantity of 10 sets. We will then be in a position to establish a selling price to the MGB owner.

So, if you are interested in possibly purchasing the AVT system let me know. Also, if you have any other comments I will be glad to hear them.