The design of a desmodromic valve mechanism for an internal combustion engine

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ABSTRACT 1

In conventional internal combustion engines, gas flow in and

out of the cylinders is controlled by poppet valves actuated by

a disc cam (usually through some linkage). A spring is used which

compresses as the valve lifts off its seat. This spring provides

the force which overcomes the inertia of the valve train during

that part of the cycle when the acceleration reverses direction

and the inertia forces tend to make the valve train lose contact

with the cam surface. This spring provides the force which

over-comes the inertia of the valve train during that part of the cycle

when the acceleration reverses direction and the inertia forces

tend to make the valve train lose contact with the cam surface.

This spring, then, controls the valve motion during the part of

the lift cycle whentthe valve is decelerating. However, vibrational

problems are inherent in systems containing springs, and resonating

and surging of the spring can be a major problem with this type

of design, sometimes leading to the breakage of the spring and

failure of the system.

In contrast to a valve system which is partially spring

con-trolled, a completely desmodromic valve mechanism may be defined

as a system which provides valve actuation which is not

spring-controlled during any part of the lift cycle. A fully desmodromic

system does not contain any springs. The term "desmodromicl is

usually associated with poppet valves, but if it is enlarged to

include any type of valves, then a sleeve valve engine has

desmo-dromic valve actuation.

A semi-desmodromic valve mechanism may be defined as one

which contains a precompressed spring which is not further

compress-ed as the valve lifts off.iits seat. This spring is merely used

to overcome "the inertia" Qf the valve k&rswhen the acaeleration

changes sign. Since the spting does not experience cyclical

com-pression and expansion, vibrational problems are not a factor. The spring functions at constant lengU] Iless the engine speed is raised above that which was used to calculate the spring preload

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necessary to overcome the inertia forces.

A valve actuation system can also be designed which combines

the best features of full dedmodromic and conventional systems.

This type of system has many advantages under certain conditions.

This might be called a three-quarter desmodromic system.

The valve actuation system designed is meant to be run as a

three-quarter desmodromic system, although it can easily be converted

to a semi-desmodromic system. It is designed for a standard C F R

eggtna, using as many standard parts as could be retained.

The design outlinedhere is for the standard C F R test engine.

The standard crankcase camshaft mounting id retained and the valves

are actuated by means of pushrods and levers. An engine speed of

5000R RPM was used for calculation of accelerations and stresses.

Initially, it was hoped that the design could be completed

in time to build and test the system, but this proved impossible

to accomplish Within the alloted time. Therefore, the project was

condidered as a design thesis only, with the hope that it might be built and tested as a separate thesis at some future date.

Thesis Supervisor:- ---

7---Prof. A. R. Rogowski

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INDEX SUBJECT 1. Title Page 2. Abstract 3. Index 4. Acknowledgements 5. Introduction 6. Design Considerations

7. Comparitive Loads on the Camshaft for

Desmodromic and Convential Systems

Design Conclusions General Conclusions Figures 1 through 11 Sample Calculations Figures 31 and S2 Appendix Figures Al through A7 Design Blueprints PAGES 1-2

3

4 5-7

8-16

17-19

8.

9.

10. 11. 12. 13. 14. 15. 20 21 22-32 33-42 43-44 Al -A4

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I wish to thank Professor A. R. Rogowski of Sloan Laboratory for his advice on theoretical and practical aspects of the design.

Mr J. Collagero's intimate knowledge of the C F R engine and

available special components proved invaluable. I also wish to thank Mr. R. J. Raymond and Mr. P. Wassmouth for their advice and assistance.

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'5

INTRODUCTION

In conventional practice, the valve actuation of an internal

combustion &ngine is accomplished by means of a disk cam which

con-trols the valve motion through various linkages (except in some

engines where the cam actuates the valve directly). The

recip-rocating parts are held in contadt with the cam by means of a spring

which is compressed as the valve is lifted off its seat. The valve

gear must be decelerated during the period of maximum lift when the

acceleration changes sign. During this part of the lift cycle, the

spring must be relied upon to provide enough force to balance the

inertia force due to the valve gear, so as to keep the linkage in

contact with the gear at all times. Therefore, during this part

of the lift cycle, the valve motion is spring controlled.

Because the spring is subjected to cyclical compression,

vibrational problems often aride. The spring, for instance, may

be subjected to a forcing frequency equal to the natural frequency

of the spring mass system. In this case, the spring often fails

to perform its function and the valve gear loses contact with the

cam surface. Usually, the natural frequency of the system is far

above the maximum driving frequency, but the driving motion is an

irregular one, and, as such, must be represented by a Fourier

series. If one of the lower harmonics of the driving frequency is

equal to the natural frequency, then trouble will result. The

kafmov CS

higher frequencyg are small in amplitude and generally are not

significant.

The spring may also fail because of surging. Surging is

essentially caused by the compression wave which runs up and down

the spring when it is subjected to a sudden displacement. This

wave is caused by the inertia lag of the center of the spring.

In extreme cases it can cause the spring to become overstressed a and break.

Valve springs also have another limitation. The maximum force

which can be exerted by the valve spring is fixed by the spring rate

and the maximum total compression. Valve accelerations, however,

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6

rise as the square of the engine speed. Therefore, if the engine

is run above the speed where the spring force is just sufficient

to balance the total inertia force, the valve will tend to lift

be-yond the maximum lift provided bqr the camand smash against the cam

surface on the way down, causing shock loads and scuffing. If the

is run still faster, the valves may not close at all, but merely

oscillate in an open position. Since the burning gases can blow

out through both intake and exhaust ports under this condition, the

valves may be melted or a fire may start in the intake manifold, or the valves may bounce against the piston heads.

Valve bounce, however, can be plainly heard and is a warning

to slow the engine down. For this reason, valve springs are usually

designed to bounce below the speed at which some of the other

re-ciprocating engine parts, such as the crankshaft or connecting rods,

are liable to fail. Valve bounce usually produces insignificant

damage if the engine speed is reduced immediately, but a broken

crank-shaft is a broken crankcrank-shaft.

At times, when valve springs have become a limiting factor in engine development, designers have tried to design valve mechahisms

without valve springs, or mechanisms in which the spring troubles

could be eliminated or minimized. These are so-called "desmodromic"

valve mechanisms. Pictures and a brief description of some of these

are given in the Appendix.

A fully desmodromic valve mechanism may be defined as a

spring-less valve actuating mechanism. In this sort of design, the valve

gear will follow the cam profile closely at any speed up to the speed where one of the components fails through overstressing. This sort

of system requires extremely accurate machining and close adjustment. Also, it is very difficult to allow for thermal expansion in this

sort of design. _{Its advantages are that it completely eliminates}

spring troubles and usually puts less of a load on the valve train than a spring-type valve systemfollowing the same lift curve at

the same engine speed. _{The most sucessfl desmodromic system built}

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7

in the Appendix).

A semi-desmodromic system may be defined as one which contains

a precompressed spring which is not compressed when the valve

is actuated at less than the designed maximum rate. The Ducatti

system

(

see Fig.A-7) can operate in this aay. If the force

developed by the precompressed spring is sufficient to balance

the inertia force of the components which are driven by the spring,

then the spring will neither extend nor compress and no vibrational

problems will come about. The advantages of this system are that

the spring allows for thermal expansion and permits larger

ma-chining tolerances. The main disadvantage is that the average

load on the valve train is higher because of the constant spring

force and the added mass of the spring and retainers.

A third type of desmodromic system might be called

three-quarter desmodromic. This type of system can be explained by

again looking at the Ducatti system. A three-quarter desmodromic

system cotains a spring as in the Ducatti system, but the spring A

has a very low spring constant. When the valve is on its seat,

there is clearance between the opening cam and the valve stem.

The spring is being held compressed by the forked end of the bell

crank and provides a small force to seat the valve. Now, the cam

surfaces can be designed so that as lift starts, the opening cam

takes up almost all of the clearance between the flanges of the

spring retainers befbre the closing cam surface starts to drop

away. If the spring force is weak enough, the inertia force at

any engine speed in the operating range of the motor will keep

the spring fully compressed at all times. Some vibration is

bound to occur in the spring, but it should not be too serious

if the total clearance between the opening and closing cams is

small. This system has most of the advantages of the fully

desmodromic and semi-desmodromic systems. However, the fact that the spring is subjected to cyclical loadihg may lead to the very sort of spring troubles that the system is supposed to eliminate.

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B

DESIGN CONSIDERATIONS

In attacking any design problem it is first necessary to

de-termine the considerations which limit the design. In this case

there were four general considerations:

1. The valve mechanism had to be for a C F R test engine

with standard crankcase. No modifications could be

made to the engine which would prevent its being

re-assembled in standard form.

2. The design should incorporate as many standard parts as possible, for financial reasons.

3. The design should be simp1qr#and-etsy and fairly

inexpensive to make and maintain.

4. The valve mechanism should be an improvement in som way

over the standard CFR valve mechanism.

Working within these considerations, and making compromises

where necessary, a design was evolved which is felt to have

enough advantages to justify its construction. The easiest way

to describe the design is to outline the principal parts and

point out the parameters which influenced the selection or design

of each part, after which the general characteristics of the

whole design can be considered.

1. Engine Crankcase: The engine crankcase used as the base

of the design is the standard crankcase for a CFR test

engine (made by War.kesha Motor Co.). Unfortunately, the

design of the crankcase prevented any practical means of

providing an arrangement to drive an overhead camshaft.

This was a big problem because an overhead camshaft would. have reduced the reciprocating mass of the system by a

factor of three, or so. However, such a drive could only

be obtained by some sort of bevel gear and shaft arrangement from the nose of the crankshaft, and this would entail new

castings and considerable engine modification. Therefore,

the standard camshaft position in the crankchs&tis retained

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9

compared to the standard two, the crankcase must be cut

away somewhat to allow installation of a guide block

for the four cam followers. A similar block with two

holes with standard spacing will permit the crankcase to

be used with standard valve gear and camshaft. This is

the only crankcase modification.

2. Camshaft: This is the most complicated component in the

design. Theccamshaft is exactly like the standard CFR

camshaft except that there are four cam surfaces instead

of two.

2a. Cam Layout: Because there is only a limited

length available for these four cam surfaces

(

4 1/8" at the most), it was desirable to have

the cam followers have as small a face diameter

as possible. Also, it was desirable to use the

standard CFR cam followers if at all possible.

However, these have a face diameter of 1 3/16"

and four of them would not fit in the available

space. Now the minimum face diameter for any

cam is a function of the cam profile and the

width of the cam. It had already been decided

to use an assymmetrical mushroom cam profile in order to allow the use of flat-fpeed followers on

all four cam lobes. dith this type of cam profile

the minimum face radius increases with the widths of the cam lobe and with maximum lift at the cam

surfaces. _{Now the tentative maximum lift for both}

valves was .36", and it was decided to retain the standard lobe width of 11/16" for the opening cams and to use a width of 3/4" for the closing cams. In order to reduce the minimum face diameter of the cam followers, it was decided to reduce the cam and produce the maximum lift of .36 at the valve by using a ratio of 2:1 on the opening rocker arms and

3:1 on the closing rocker arms. The lobe widths r

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10

surfaces, By this means, it is possible to

use standard CFR cam followers with the face

diameter ground down to 1" for the opening cam

lobes and 7/8" for the closing cam lobes. The

reduction in lift was alsom necessary for another

reason. The cravnkshaft balance weights and the

bolt heads on the connecting rods pass very close

to the camshaft during part of their cycle. With

only two cam surfaces, the cam lobes can be

positioned so as to avoid any metal-to-metal contact.

With four cam lobes to fit into the same space,

however, the problem becomes much greater.

Measurements on layout drawings of the engine and

actual measurements on one of the CFR test engines

in Sloan Laboratory were necessary to insure that

the cam lobes would not interfere with crankshaft

rotation. These measurements were also used to

determine the maximum permissible zero-circle radius

of the closing cam lobes.

Because of the restricted space available, the

cam lobes are not situated symmetrically about the

cylinder centerline. Moreover, not all the space

available could be used, as it was desirable to

keep the centerline of the pushrods well inside

the centerlines of the studs which locate the

cylinder head on the 'cylinder. Because of this,

the centerlines 0f the closing cam followers could

only be offset 1/32" from the centerlines of the

closing cam lobes, although it was possible to offset

the opening cam followers 1/8".

2b. Cam Profiles and Timing: As already stated, it was

decided to use flat cam followers. This required

that the cam surfaces should at no point be

concave. This was easily ensured for the opening

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cams profiles, however, had to be calculated

from the opening cam profiles allowing for the

rocker arm ratios. A very accurate 10x gcale

drawing was then made and the surface was

constructed to an accuracy of about plus or minus

one-hundredth of an inch to dtermine whether the

surface could be used with a flat follower, or

whether it would be necessary to use a roller

follower. In scale, this was plus or minus one

thousandth of an inch, so this method provided a

very accurate picture of the cam profile. Luckily,

the surface proved to be convex at all points, so

a flat follower could be used.

The cam profiles used are rather unusual in

that they are not symmetrical about the point

of maximum lift. This can be 3een most

dramatically in a plot of the valve(accelerations

(see Fig. 1). -hese valve accelerations are

calculated at 5000 RPM (engine) which is a

piston speed of 3750 fpm for "the CFR engine

(4:500"

stroke). Professor A.R. Rogowski of Sloan

Automotive Laboratory, M.I.T., told me that Sloan Lab has run CFR engines at 4500 RPM (3370 fpm piston speed) so the engine is capable of at

least that engine speed.

Figure 2 shows the lift of the valves versus

crankshaft angle- _{From these two plots, it can}

be easily seen that the lift pattern has a "fast"

side and atslow" side. _{The "fast" side moves the}

distance of maximum lift in 120 crankshaft degrees, the"slow" side moves the same distance in 150

crankshaft degrees. '1V

The exhaust valve lift curve has a fast

opening side and a slow closing side. This is done

in an attempt to open the exhaust valve wide during the period when the pressure in the cylinder is

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12 blowdown may be more complete and pumping work

re-duced during the exhaust stroke. The valve is closed

slowly because it is felt that there is no real point in closing it faster.

The intake valve lift curve has a slow opening

side and a fast closing side. If the piston is

re-garded as a pump during the intake stroke and exhaust strokes, then maximum valve lift( and therefore max-imum flow area) should probably come at the point of

maximum piston velocity. In this design, however,

the maximuk flow area of the exhaust valve occurs

before maximum piston velocity in order to make best

use of the pressure in the cylinder when the exhaust valve opens. The point of maximum flow area for

the intake valve comes after maximum piston velocity

(see Fig. 3). Now, when running at high speed

the flow velocity will lag behind the piston

velocity, anyway, because of inertia. Since

this valve system is primarily meant fer

high speed operation, it is desirable that maximum flow area for the intake valve should occur after maximam piston velocity.

In reality, inertia and pressure wave effects can be so arranj2ed that it would be desirable to have the inlet valve wide open during its whole

cycle. This, of course, requires infinite

acceleration and is therefore impossible from a

kinematic standpoint. _{The intake cams, however,}

could have had a fast opening side (120 crank degrees), a dwell of 30 crank degrees, and a fast closing side (120 crank degrees) without increasing

the duration of the lift. This would substantially

increase the maximum total load On the camshaft. If both intake and exhaust cams had had this profile, the maximum total camshaft loading would have been

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very much larger, since this maximum occurs

during the period of valve overlap.

Valve timing must always be a compromise. The

large amount of valve overlap and large duration

period which allows adequate breathing at high engine speed permits flow reversal at low engine

speeds. Racing emgines generally have high volumetric

efficiency at high engine speed, but their volumetric

efficiency at low engine speeds is very poor due to the valve timing.

Desmodromic valve mechanisms have the

capability of reducing valve overlap and duration

period without any sacrifice in average flow area

by substantially increasing the valve accelera'tions.

The Mercedes-Benz desmodromic valve design operated

+ke eK9)IAe.

at peak valve accelerations of 1500 "g" andAhad a very flat torque carve over quite a range of engine speed with maximum volumetric efficiencies over 1000.

The valve timing used on the camshaft in this design is moderate in an attempt to make the engine

fairly flexible. _{V&lve accelerations are not very}

high because of the large reciprocating mass involved.

5. Push Rods: _{The standard CFR push rods are lengths of}

5/16" diameter round, steel rod w'ith the ends rounded.

Using'Euler criteria, their critical load (the load at

for

which they buckle) is about 890 lb & 6ngth-Appdoximately that required for the design. Now the peak load on the

opening pushrods is 7251bs. and the peak load on the closing

pushrods is 975 lbs. _{These pushrods are therefore inadequate}

for this design. _{1ew pushrods were therefore designed using}

Shelby hard-drawn steel tubing (.,75" Q.D., .049" wall

thickness) with pressed-in ends. These have critical loads of 1310 lbs. for the opening pushrods and 1500 lbs. for the closing pushrods (which are slightly shorter).

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4. Opening Rocker Forms: These rocker arms were originally designed to be made of steel, but were redesigned to be

made of duralumin. Steel inserts are pressed into the

end which contacts the valve stem to minimise waar. At

peak load, using a simplified model, calculated

maximum deflectionat the valveis .0072". The model

used is decidedly pessimistic and the actual maximum deflection

under peak load should be a small fraction of the calculated

valve. Liaximum stress occurs around the hole for the

pivot shaft brushing. Using a stress concentration

factor of 3, the maximum fiber stress at this point is

about 18,000 psi, well below the yield point for the

best aluminum alloys. The opening rocker arms multiply

the opening cam lift by a factor of twoas stated before. An analysis was made of the Change in lift ratio due

to pushod angularity and change in contact radius due to

rocker arm rotation and valve motion for both opening

and closing rocker armsto determine whether this

factor should be figured into the cam profiles. However,

it was found that the total effect was not as large, percentage-wise, as the machining tolerance for the cam

profiles.

5.Closing Rocker Arms: these parts are the sore point of

the whole design. They have to be rather long because

of the geometry of the engine, and the forked end concentrates

quite a bit of mass ataa large distance from the axis of

rotation. As a result, they have a large moment of inertia.

These parts were also originally designed to be made of

steel and were redesigned to be made of duralumin with

steel inserts at the contact points. Again using a

simplified model, maximum deflection at the valve was

figured to be .041". _{The model used here is very pessimistic,}

but actual deflection may be about .(SZ5". If deflection

in practice proves to be larger than this, new parts should be made of steel with a slight increase in total reciprocating

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mass. The closing rockers multiply the closing cam lift

by a factor of 3, as stated before. A sketch of the rocker

arm assembly is shown in Fig. 5.

6. Valves: the valves used are made by Thompson Products. They

have a face diameter of 1.32" (therefore the lift of .36" is

27% of the valve head diameter). The, stem diameter of these

valves is the same as he standard C F R valves, so standard valve guides may be used. The overall length of the valve is 6:' 3/16". This particular valve is used becausgt is longer than the standard C F R valvelbut weighs the same. The, extra length was necessary to provide clearance for

the closing rocker arm as it moves down.

7. Cylinder Head: the cylinder head used in the setup is a

standard removable C F R cylinder head. The only modifications necessary are to remove all obstructions on the top face

of the head. This involves moving the coolant outlet pipe to the side of the head, plugging the hole, and machining the valve guides off flush with the top surface.

8. Cylinder: the cylinder used in this design is a fabricated

cylinder with standard C F R bore (3.250"), originally used in some air compressor tests. The coolant outlet pipe must be moved because it interferes with the push-rodqland the base cut out somewhat to accommodate the guide block for the cam follower. These are the only modifications.

9. Springs: when the system is run as a semi-desmodromic system,

a spring has been calculated which has a free length of 1.25" (working length about 1"t). The spring constant is 575 lb/in.

The spring is precompressed to give a force of 150 lbs, which is sufficient to balance the inertia forces at 5000 R P M engine speed with as excess of 12Ibs. to take care of friction. _{If the friction is greater, the spring may} be further compressed initially.

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If it is desired to run the system as a three-quarter

desmodromic system, the above springsshould be replaced with

a standard spring which can be obtained from most spring

manufacturers. This spring should have a free length of

2" and should be able to be fully compressed without

set-ting((compression springs ardften designed on this basis).

The spring constant should be around 20 lb/in. At a

work-ing length of 1", this will give 20 lb. of force to seat

the valve. When running the system as a three-quarter

desmodromic system, the bottom spring retainers should

be replaced with the alternate ones which have been

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17

COMPARATIVE LOADS ON THE CAMSHAFT

FOR DESMODROMIC AND CONVENTIONAL VALVE SYSTEMS

The camshaft is subjected to higher total loads than any other

valve train component. The camshaft loads are therefore indicative

of the relative load that any of the components has to carry.

Therefore, the total camshaft loads have been calculated and plotted

against camshaft rotation for the system used as a semi-desmodromic

system and as a three-quarter desmodromic system. In addition, the

camshaft loads have been calculated for a conventional system using

springs and following the same lift curve. In practice, this can

be easily done with the designed system by removing the closing

rocker arms, pushrods, and cam followers, and installing springs

to close the valves. Th spring necessary for this job should have

a freelength of 2 1/4". The initial compression with the valve

in the closed position should be 1/4"; a hardened spacer about 1/4"

thick must beplaced betweentthe spring and the cylinder head to

achieve this static load and prevent cylinder head wear. The spring

constant is 264 lb/in.

For purposes of comparison the cam loadings have also been

cal-culated and plotted for a fully desmodromic system of the

Mercedes-Benz type with an overhead camshaft. The same lift curve is followed.

In plotting the total camshaft loads, the forces have been

assumed to be all in the same direction. This is true in all cases

except for the fully desmodromic system where the' forces on the closing cam lobes are at about 90 degrees to the forces on the

opening cam lobes. For the sake of simplicity and consistency,

however, the loads on the camshaft of the fully desmodromic system have been piuAied as if they were in the same direction.

Figures 6,7,8,and 9, show the loads on the separate cam lobes when the system is being run as a semi-desmodromic system at an

engine speed of 5000 R P M. The loads on the individual cam lobes

when the system is run as a three-quarter desmodromic system are

5 kowti

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1 3

(spring) load (40 lbs on the opening cam lobes, 60 6n the closing

cam lobes. Atlthe points where the inertia loads are negative,

there is no load on that particular lobe.

Figure 10 shows camshaft loads (at 5000R P M engine speed) as a function of camshaft angle for the designed valve mechanism operated as a semi-desmodromic system and for the same mechanism using valve springs to effect closure.

The semi-desmodromic system has much higher camshaft loads

than the conventional system. This is principally due to the

high reciprocating mass which requires that the spring preload be rather large.

The peak load on the camshaft with the semidesmodromic PYstem

is about 1700 lbs. If the camshaft is treated as a round beam

4 inches long and 1 1/8 inch in diameter with built in ends and with a 1700 lb load at the center, the maximum deflection is

.00024 inches, and the maximum fiber stress is 8200 psi. In the actual piece, the stress will be greater by a factor of 4 or so at the junction of the shaft and the bearing flange. Also, the fact that the shaft is rotating makes the situation somewhat worse as do the varying torsional loads in both directions at each cam

lobe. _{However, the model, which has all the load concentrated}

at the center of the beam, tends to give deflections and stresses which are larger than those which would lbe obtained if the loads were distributed as they are on the actual camshaft.

The actual camshaft deflections and stresses, taking all of the factors into account, could be calculated only by setting up the problem and running it through a computer, and even this would merely give approximate results.

Figure 11' shows camshaft loads (at 5000 R P M engine speed) for

the designed valve mechanism when used as a three-quarter

desmo-dromic system. _{Also plotted are the loads for a fully desmodromic}

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calculations for this latter are-based on estimated reciprocating

mass for a system of this type mounted on a C F R engine, using

valves having the same weight, and following the same lift curve.

It, therefore , is a fairly accurate picture of the relative size

of the loads involved.

Figures 10 and 11 are drawn to the same scale, so direct

com-parisons cai be made. The curves are plotted on two separate graphs

only for the sake of clarity.

Comparison of the four curves shows that the semi-desmodromic

system is definitely inferior to the conventional system in this

instance, unless spring problems are present and can not be

elim-inated. The three-quarter desmodromic system, on the other hand,

has lower peak loads, despite all the added reciprocating mass. The curve of the fully desmodromic system is very much lower

than any of the others, and demonstrates the advantages of

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20 DESIGN CONCLUSIONS

The designed valve mechanismican be operated as a

semi-desmo-dromic system. Indeed, all the stress and deflection calculations

are based on the peak loads encountered in hhis sort of operation,

to ensure that it would be possible to run in this manner.

However, if this system is built, it is advised that it be run as a three-quarter desmodromic because of the lower loads (and

there-fore lower wear)encountered.

If the system is run as a semi-desmodromic system, the spring otAsta:Wt

should be carefullyh hecked and the initial compression set carefully

in order to apply the correct preload.

If the system is run as a three-quarter desmodromic system, the

bottom retainer which has been designed for this purpose'should be

installed. T'he clearance between the opening rocker arm and the

valve button should be set at about .015" with the valve on its seat. The engine should then be turned over by hand until the valve is

lifted slightly off its seat. The clearance between the flanges

of the two spring retainers should then be set at .005". The engine

should be turned over completely a few times to check for interference

of any sort. It can then be motored at a range of speeds to make

sure that the spring does not surge.

In this case, spring surge may be caused by the acceleration of the spring, not by the relatively small displacement which i

it undergoes. Under these conditions, the vibrating mass is equal

to one half the mass of the spring, and the effective spring constant

seen by this mass is twice the spring constant of the spring itself.

For the spring described earlier, weight is .046 lb. and spring

constant is 20 lb/in. The natural frequency of surging is therefore

7800 cycles per minute. _{Since maximum camshaft speed is 2500 R P M,}

only the fourth and higher harmonics of the motion have any chance

of exciting resonance. The amplitude of these harmonics is usually

small; however, if surging does occur, the spring should be changed and either a lighter spring, or one with a larger spring constant substituted.

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GENERAL CONCLUSIONS

As stated before, it is believed that desmodromic valve systems

will be seen on production internal combustion engines in the future,

probably on the optional high performance engines which most

manu-facturers offer, and almost certainly on production sports car engines.

These desmodromic valve systems will probably be of the three-quarter

type. Very likely, the design will have both opening and closing

cams on an overhead camshaft, although it would also be possible

(and more desirable, on a production engine) to have two cams: an opening cam in the standard position in the crankcase, and a closing

cam in an overhead location, closing the valves by means of levers

(23)

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(24)

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(25)

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(26)

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