An Examination of the Torsional Vibration Characteristicsof the Allison V-1710 and Rolls-Royce Merlin Aircraft Engines
by Robert J. Raymond and Daniel D. Whitney
June 2016

PrefaceBy 1940, most of the high output radial aircraft
engines were utilizing tuned pendulum dampers
to minimize torsional vibrations in the crankshaftpropeller system. The situation in the in-line highoutput configuration was less consistent. While anumber of earlier engines had adopted various
types of dampers (mainly of the friction type) by
1940 all but one engine, the Allison V-1710, were
without dampers of any kind. Rolls-Royce,
Daimler-Benz and Junkers, all with liquid cooled
V-12s and operating at comparably high outputs,
were damper free. The Rolls-Royce Griffon, underdevelopment but not yet in service, never
employed a damper.
This paper is an attempt to explain why the
Allison engine was unique in this respect. To doso, I am using the Merlin to compare the torsional
characteristics of the two engines since they areclose in displacement with similar bore/stroke
ratios. There is nothing in their designs that could
lead one to think their vibration damping characteristics might be different. If anything, the larger
main and connecting rod bearings in the Allison
would indicate more damping and, as I haveshown elsewhere, the friction mep of the two
engines was probably very similar.
The data needed to carry out this analysis are
not entirely complete but more information isavailable than for any other similar engines. In
their published work Allison does not present test
data that indicates a need for dampers in any documents available to me nor have I found actual
torsiograph data that indicates they are necessary
to stay within the limits specified by Army-Navystandards.

Introduction
The subject of torsional vibration in piston
engines is difficult to digest in one sitting. Most

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mechanical engineers are not exposed to the subject either in school or in their work environment
but, of course, are familiar with the fundamentalsof vibration. The numerous terms — nodes,
modes, orders (major and minor), critical speeds,
etc. — are confusing when first encountered. I
will try to explicate as we go along but the reader
may want to look at my paper on the Liberty-12
available on the AEHS web site where the analysis is more detailed than it will be here.
The crankshaft is like a violin string; it canhave many modes of vibration, the simplest in thecase of the violin being a half wave length with anode at either end. Superimposed on this fundamental mode are multiple higher frequency
modes. Twisting rather than lateral deflection
characterizes torsional vibration in a crankshaft
and the first two modes are the most important.
The first mode can be thought of as the engine
assembly vibrating against the propeller with asingle node located in the propeller shaft and the
maximum angular deflection at the rear of the
crankshaft. The second mode is at a much higher
frequency and has two nodes, one in the center ofthe crankshaft and a second in the propeller shaft
near the propeller. Maximum deflections are usually at the rear and front of the crankshaft or the
gearbox, depending on the relative stiffness of thecoupling between the crank and the gearbox. Thismode can be thought of as the front and rear
halves of the crank vibrating against each other.
Determining the shape of these modes andtheir associated natural frequencies is the firststep in any analysis of torsional vibration in anengine’s rotating assembly. This is accomplishedby replacing the crank, connecting rods and pistons with a series of flywheels and shafts that represent the inertia and stiffness of the various elements in the system. Through engine tests withtorsiographs, which measure the frequency andamplitude of vibration of a crankshaft, and static
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1stiffness tests on crankshafts this technique wasdeveloped to give satisfactory results. These models are called mass-elastic diagrams and Figure 1
is an example. The natural frequency is calculated
by giving the first inertia a one-radian deflectionand estimating a frequency. If the estimate isincorrect there will be a remainder torque at thelast inertia and the procedure is repeated until theremainder is zero. This is known as the Holzer
method.
Once the two natural frequencies are established the next task is to investigate the forces that
excite the vibrations in the crankshaft. It is obvious that the torque due to pressure and inertia
varies dramatically over the 720-degree cycle foreach cylinder, with about a 20% variation in output torque between the last pair of cylinders andthe propeller in a V-12 engine and much more
drastic variations as you go back toward the rearof the engine. The magnitude and amplitude ofthese fluctuations determine the stress level andestablish the fatigue limit of the crankshaft. To
these values one must add the stress induced by
the vibration of the crankshaft. It is intuitively
obvious that there are six equally spaced strong
pulses in one revolution of the crankshaft so one
would expect that an engine speed corresponding
to one sixth of a crankshaft’s natural frequency
might be problematical, which in fact it is. The
sixth is a major order of excitation in a V-12 engine.
The magnitudes of all the other orders, major andminor, are determined by carrying out a Fourier
analysis of the torque versus crank angle curveand reducing it to an equivalent series of sinusoidal curves of varying frequency and amplitude. These are then represented by vectors which
are combined for all of the cylinders to give aresultant value or phase vector sum. The various
orders vary in frequency from that corresponding
to 1/2 engine speed in frequency increments of
1/2 to about 8, beyond which their magnitude
becomes insignificant. When all of the vectors of agiven order point in the same direction, that is amajor order — like the sixth just mentioned.
When they point in different directions, usuallysymmetrically around a center, they are called
minor orders. Because the crankshaft does not
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deflect uniformly along its length the minor
orders can be troublesome as well.
Each order of vibratory torque has a magnitude
that depends on the mean effective pressure, compression ratio, and, to a lesser degree, otherengine operating variables. This vibratory torquemultiplied by the vector sum associated with thatorder is the excitation torque for torsional vibration. If, for instance, the engine is operating at
2,500 rpm and the natural frequency of the crankshaft system is 6,250 vibrations per minute the 2½
order would excite vibrations in it. If there wereno damping in the engine the amplitude of thevibration would be infinite and the crank would
fail. At this point the designer would need toknow the damping characteristic of the engine inorder to calculate the amplitude, and hence thestress, in the crankshaft.
By the mid 1930s the technique just outlined
had been refined enough that the designer could
modify values of stiffness in the system so as to
avoid severe torsional vibration problems. He
may not have been able to predict the amplitudeof vibration without prior experience with a similar engine design but the magnitude of the various excitation torques and associated critical
speeds could be ascertained and influence thedesign process.

Analysis
As mentioned above, the first step is to construct a mass-elastic diagram for the engine to beanalyzed. In this case I have chosen two configurations of the Allison engine and one of the
Merlin. These are shown in Figures 1, 2 and 3.
Figure 1 is a composite of two mass-elastic diagrams provided by Dan Whitney for the V-1710-E(shown in Figure 3) and the twin crankV-3420
with close-coupled gearbox. I created the composite of Figure 1 to be closer to the Merlin’s configuration because I thought it would provide a better
comparison since the E version of the Allison
engine had a long, flexible extension shaft andremote gearbox. There were at least two different
couplings used by Allison to connect the

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RELATIVE
DEFLECTIONONE NODE
fn = 6,335 vpm

STATIONTWO NODE
fn = 18,235 vpm
RELATIVE
DEFLECTIONSTATIONFig 1. Mass-Elastic Diagram and Relativie Deflections for Two Modes of Vibration – Allison V-1710 Composite

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RELATIVE
DEFLECTIONONE NODE
fn = 4,870 vpm

STATIONTWO NODE
fn = 18,567 vpm
RELATIVE
DEFLECTIONSTATION

Fig. 2. Mass-Elastic Diagram and Relative Deflections for Two Modes of Vibration – Rolls-Royce Merlin

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RELATIVE
DEFLECTIONONE NODE
fn = 2,817 vpm

STATIONTWO NODE
fn = 16,417 vpm
RELATIVE
DEFLECTIONSTATION

Fig. 3. Mass-Elastic Diagram and Relative Deflections for Two Modes of Vibration – Allison V-1710-E

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5crankshaft to the gearbox in the close-coupledconfigurations and information provided by Dan
Whitney indicates that the numbers I used in
Figure 1 conform to the later coupling design.
The Merlin mass-elastic diagram (Fig. 2) camefrom Reference 1 but did not include a propeller
inertia (for reasons I will get into later) so I used
the inertia of the Allison propeller. The Merlin
diagram did not include the stiffness and inertiaof the supercharger drive and I was unable tofind those numbers. I do know from the drawings
that the Merlin drive system was much more flexible than the Allison’s and in two stage Merlins
the inertia was effectively much higher so the natural frequency of the supercharger vibrating
against the engine would be quite low.
I have chosen to ignore this lowest mode ofvibration because it has little effect on vibratorystresses in the crankshaft. In calculating the natural frequency for the one and two node vibration
modes shown in Figure 1, adding the supercharger from the mass elastic diagram for the Allison
V-1710-E, not shown in Figure 3, changes the one
node frequency by about 3% and the two node by
an insignificant amount. The relative deflections
remain unchanged.
Figures 1, 2 and 3 show the natural frequencies
and relative shaft deflections at those frequencies
for the three systems we are comparing. Note therelatively large differences in the one node frequencies of the three engines compared to the twonode. This is due to the relative flexibility of the
Merlin crank to gearbox coupling as compared tothe composite Allison. This also explains the more
gradual slope of the V-1710 composite deflection
diagram. By contrast the V-1710-E one node frequency is considerably less due to the long extension shaft and its deflection curve is almost flat
for the engine portion of the curve, which resultsin much smaller minor order resultant vectors forthe one node vibration.
With the natural frequencies and the shape ofthe deflection curve established for the three cases
the next step is to determine the phase vector
sums for the various orders of excitation torques.
The Merlin and Allison had different firing orders
but the vectors combine in the same manner so
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there is no difference in the vibration characteristics of the two engines attributable to firing order.
The only major orders in the operating range ofthese engines are the third and the sixth. The
third order vector sum is zero for a 60° V-12 andthe sixth order occurs only in the two node operating speed range so that the vectors associatedwith the back half of the crank are balanced to agreat degree by the vectors from the front half. I
examined the phase vector sums for all the orders
in the operating range for one and two nodemodes of vibration. The ones with significant
magnitude are included in Table 1.
Column 8 of Table 1 gives the resultant vector
sums for orders that result in significant vibratory
torque in the operating speed range of the three
engines.
The next step is to determine the magnitude of
the torques associated with the various orders.
Reference 2 (an Allison paper) gives values ofthese (in terms of tangential pressure) for orders½
to 6. I extrapolated their numbers to get the higher
orders. The Allison numbers were consistent withthe generalized data for spark ignition engines
and I am assuming that the Merlin’s are the sameat the same imep. I converted their numbers frompressure to the ratio of vibratory torque to mean
torque per cylinder. The mean torque is a functionof the indicated mean effective pressure, which, inthis case, follows a propeller curve up to about
2,000 rpm and 188 psi imep and, with the engine
at full throttle gradually increases to 210 psi at
3,500 rpm. This is the load curve Allison used in
its report to the Air Corps on the V-1710-E in 1939
and it contains a torsional analysis of the engine
that would come to be equipped with thehydraulic and pendulum dampers for the remaining life of all the various V-1710 models. I thought
it appropriate to use this load curve since the ratings of the Allison and Merlin were roughly thesame in that time frame, ~1,000 horsepower.

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Table 1, columns 4 and 5, show the engine
speed and imep for the various orders under consideration. For each of these conditions there is amean torque per cylinder for the imep shown(column 6) and a vibratory torque per cylinder(column 7) for the corresponding order. The product of the vibratory torque and the phase vectorsum (column 9) gives the vibratory excitation
torque in inch-pounds per radian of deflection at
the rear of the crankshaft.
Figure 4 is a plot of the excitation torques forboth modes of vibration for the three engines.

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There is not much here to explain why the
Allison was in need of dampers and the Merlin
was not, especially considering that Allison was
designing the damper system for the V-1710-E
whose amplitudes are below the Merlin’s acrossthe board. The 4½ order, two node excitations areabout equal in magnitude but the V-1710-E ispeaking about 500 rpm lower and, therefore, closer to the operating range. It’s doubtful that thiswould be more of a problem than the Merlin 1½,
one node at 3,250 rpm.

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Fig. 4 Vibratory Excitation Torque Versus Engine Speed (see Table 1 for Mean Torque)
Note that the torque is for 1 degree of deflection at the rear of the crankshaft.

At this point I decided to try and predict anactual vibratory amplitude at the rear of the
crankshaft to see how it might compare with
Army-Navy Specification No.9504 ca.1942. This
specification limits one node vibration amplitude
to ±1.5° and two node to ±0.25°. Reference 2(Allison again) gives a curve of magnification factor at resonance versus vibration frequency. If one
equates the energy input for a particular order atresonance to the energy dissipated due to damping it is possible, with the magnification factor, tocalculate the vibration amplitude. This is assuming there are no dampers in the system, only thenatural damping in the engine itself. The
Magnification factors for the relevant natural frequencies are shown in Table1, column 12. The
equivalent inertia of the engine is given in column
10 and the static deflection is given in column 11,
both of these values are used to calculate the half
amplitude of swing at the rear of the engine, theta(column 13).
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The results of this analysis are shown in
Figures 5 and 6 for the one and two node modesof vibration.
The one node case does not appear to be a
problem for either Allison configuration while theMerlin 1½ order could be considered problematical but Rolls-Royce never used a damper in thatengine. The two node case shows everything to be
below the value allowed by the A-N spec. With
the higher engine ratings to come during the war
years the V-1710-E 4½ order could be considered
a problem, but certainly not in 1940.
The construction of a Holzer table allows one
to calculate a torsional stress occurring when the
crankshaft is in free vibration at its natural frequency since torques are calculated at each stationin the mass-elastic diagram. Table 2 shows thesestresses for our three cases.

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RPM
Fig. 5. Estimated Vibration Amplitude at Rear of Crankshaft versus Engine Speed – One Node Vibration Mode (see Table 1 for Mean Torque)

RPM
Fig. 6. Estimated Vibration Amplitude at Rear of Crankshaft versus Engine Speed – Two Node Mode of Vibration (see Table 1 for Mean Torque)

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9The Merlin’s one and two node stresses arehigher than the V-1710-E values by a significant
margin. Some of this is due to the smaller diameter of the Merlin’s journals. The two node value islower also due to the lower natural frequency inthe V-1710-E two node vibration mode and thelow one node stress is due to the very little
amount of twist in the crank as seen in Figure 3.
Again, if the Merlin didn’t need dampers the V1710-E should not have needed them either.
Table 2 also illustrates why the allowed maximum half amplitude is only a quarter of a degree
for the two-node mode. The numbers in the table
are for one degree half amplitude and restricting
the amplitude to one quarter would get the vibratory stress down to the same level as for 1½
degree one node vibration.

Discussion
This section will consist of a discussion of myresults as compared to the limited amount of data
available and experience with the two engines in
air racing and in hydroplane boat racing post
WWII.
Tests conducted at Wright Field on a PackardMerlin (Reference 3) gave a 2½ order peak of
about 0.35° at station 6 in Figure 2. Using the relative deflections of Figure 2 this translates to about
0.5° at the rear of the crankshaft. My analysis
gave about 0.6° as shown in Figure 5 so I am
apparently at least in reasonable agreement. Theirpeak occurred at about 2,300 rpm while mine
occurred at 1,950. This same Army report notes a

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7½ order at 2,700 rpm but does not give a magnitude. My analysis gave a 7½ at about 2,500 rpm.
An interesting comment in the Army paper is thatthey lacked a mass-elastic diagram for the Merlin
and planned to carry out such an analysis. One
would assume that if Packard had such an analysis they would gladly have supplied it and that
Rolls-Royce would have supplied it to them aspart of the licensing arrangements if torsional
vibrations were a significant problem. Rememberthat this was in mid 1942. A comment I received
from Dave Piggot of the Rolls-Royce Heritage
Trust on this subject may shed some light on this:“In all the time I have spent in our archive, I havenever come across any reports of some substanceon the subject of torsional vibration on either the
Merlin or Griffon”.
Figure 7 shows the results of another torsional
test on a Merlin. This was carried out at the Royal
Aircraft Establishment in 1943 (Reference 1) andis the source of the mass-elastic diagram in Figure
2. I mentioned that no propeller inertia wasincluded and that I had used the Allison propeller’s inertia. Figure 7 is similar to Figures 5 and6 except it is in terms of torque rather than angular deflection at the rear of the crankshaft. The
torque in this case was measured between stations 6 and 7 in Figure 2. Note that Figure 7
shows the same orders of vibration occurring at anumber of engine speeds and even a third order,
which should not occur. This is due to the factthat then modern aluminum variable-pitch propellers did not behave as a solid inertia as in our
mass-elastic diagram, but were subject to various
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10Fig. 7 (from Reference 1)

bending and flapping modes of vibration that
could excite vibration in the crankshaft. In essence
the propeller should be thought of as a branchedsystem of distributed masses subject to bending
and twisting. Apparently the three bladed prop
caused the third order shown in Figure 7. The 2½
order has peaks at ~1,350, 2,000, and 2,250 rpm
whereas my peak as shown in Figure 5 is at 1,950rpm. Of the three their 2,000 rpm gave the highest
amplitude and the resonant frequency recorded
was 5,000 vpm, close to my value of 4,870.
The two node orders in Figure 7 (4½ and 7½)
are almost non-existent in magnitude and theauthors concluded that “judging from the corresponding stresses in the crankshaft, the twonoded mode of crankshaft torsional vibration
appears to be relatively unimportant but it has
been found that this mode of vibration may giverise to high stresses in the propeller blades”. My
analysis as shown in Figure 6 gives a 7½ at 2,500rpm versus 2,500 rpm in Figure 7 and a 4½ order
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at 4,100 rpm versus 3,000 rpm in Figure 7. They
may have been picking up the flank of that orderor it could be that the propeller had shifted thenatural frequency as it did for the one node cases.
Runs similar to those shown in Figure 7 withchanges in the pitch setting of the propeller gavevery different results to those shown in Figure 7.
The largest 2½ order occurred at 2,100 rpm with ablade pitch of 29° and a natural frequency of 5,200vpm.
As an additional reality check the amplitude ofvibrational torque at station 6 (Fig. 2) from Figure
7 is about 9,500 inch-pounds. Using the Holzer
table for the Merlin and the vibrational amplitude
from Figure 5, 0.6°, I calculate a vibratory torqueof 9,000 inch-pounds. The analysis seems to fit
what data I have for the Merlin engine.
Unfortunately I lack any comparable data fromthe Allison engine. The Army report mentioned
previously, Reference 4, was the source of the
S.A.E. paper, Reference 2. Neither document
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11contains the results of a torsiograph test that
could explain their need for a damper to protect
the crankshaft. There is a final figure in the S.A.E.
paper that shows excitation torques and theresults of a torsiograph test with and without adamper. It is not stated whether it is one damperor both that is being applied and the magnitudes
and frequencies at which the excitation torques
occur do not correspond to the results for the V1710-E in Reference 4.
The only evidence I’ve seen for the need ofdampers in the Allison engine is Allison’s statements to that effect. Reference 5, Dan Whitney’sbook, sites a crankshaft failure at a location thatwould correspond to a maximum two node stressarea. The actual failure was attributed to a manufacturing defect but the pendulum dampers were
installed as a fix.
At this point I will introduce the damper construction used in the later model Allison engines.
Some earlier models employed a friction typedamper at the output gear. Allison’s reason given
for abandoning this approach was that it wasineffective for damping two node vibrations
where amplitudes are small. Allison’s system isshown in Figures 8 and 9.
The hydraulic damper protects the supercharger drive and dampens the one node vibration. The
pendulum dampers are tuned for the two node
4½ and 7½ orders, 3 weights for each order.
According to Dan Whitney the 4½ order weights
were removed in the later V-1710-G engines,
which, given the results of my analysis, is reallypuzzling when you look at Figure 6. I’m assuming here that my composite Allison is close to the
G model. Since the pendulum dampers were
designed originally for the E model, perhaps thelower amplitude and higher critical speed of mycomposite engine could justify removing the 4½
damper even though the G model was rated at ahigher speed than the E model. On the other
hand, the engine ratings were much higher for the“G” model so the excitation torques and hence thevibration amplitudes would have been muchhigher than those shown in Figures 4 and 6.
To summarize, my analysis of the Merlin seems
to agree with the few test results I can check it
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against. The analysis for the Allison E modelshows much more modest torsional activity thanfor the Merlin yet the Allison was equipped with
first and second mode dampers and the Merlin
had none. The only factor that can explain this is
if the Allison engine had much less natural damping than the Merlin. As I mentioned before thisseems very unlikely due to the larger main andcrankpin bearings in the Allison and the likely
similar friction mean effective pressure (seeReference 6).
The experience with both of these engines post
WW2 in aircraft and hydroplane boat racing provides some additional perspective on this question. Both applications resulted in engine speeds
well above their rated speeds in military aircraft,routinely into the mid 4,000 rpm range. In theboat application the Merlin failed crankshafts
until additional counterweights were added toreduce main bearing loads, particularly the center
main. Without the additional counterweighting
the center main bearing cap bolts would yield andstretch reducing support for the crank. Once theweights were added no further crankshaft failures
were experienced. The stock supercharger drive
system in the Merlin had a life of about 15 minutes, apparently due to rapid engine speed
changes when the boat’s propeller came in andout of the water failing the quill shaft and overrunning clutch. The fix for this was a stiffer andstronger quill shaft. My source for this information was Dixon Smith who had personal experience in this area and is a practicing mechanicalengineer. He does not know of anyone installing
dampers in Merlins but Dan Whitney has apparently heard of at least one. It should be noted thatthe torsional characteristics of both engines would
have been changed significantly due to the differences in the propeller inertia and the change in
the gear box and coupling to accommodate aspeed increase in boat racing applications. I doubtthat these changes would have changed the twonode mode of vibration very much.
Experience with the Allison engine was somewhat different. There apparently were no main
bearing cap bolt stretching problems, probably
due to the more heavily counter-weighted design
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12Fig. 8. Allison Gear Train Showing Dynamic and Hydraulic Dampers

Fig 9. Allison Hydraulic (left) and Dynamic (right) Dampers
1 = Nut; 2 = Weight, Small; 3 = Pins; 4 = Hub; 5 = Weight,
Small; 6 = Crankshaft Flange; 7 = Pin Retainer; 8 = Bolt

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Fig. 10. Allison Hydraulic Damper Operation (from Reference 5)

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The analysis is based on ca.1940 ratings andthese two engines ultimately were rated at roughly twice that power. If we double the vibratory
excitation torques and look at Figures 5 and 6 we
see that we are still at or near the Army-Navy
specs at speeds under 3,000 rpm. The damper
weights in the Allison were the same in 1945 asthey were in 1939, despite the increase in ratings,
which implies they had sufficient amplitude to
generate an adequate reaction couple to theincreased vibratory excitation torque.
Dan Whitney has suggested the most plausible
scenario for the adoption of dampers in the
V-1710. The V-1710-E was designed for the BellXP-39 aircraft where the engine was locatedbehind the pilot and a cannon fired through thegearbox and propeller shaft. This was the reason
for the long extension shaft (Fig. 10). Testing
revealed misfiring above about 2,600 rpm. This
was traced to the magneto and attributed to torsional vibrations exciting a resonance in thecamshaft and magneto drive system, which was
driven off the rear of the crankshaft. This apparently developed into somewhat of a panic situaConclusion
tion and, it seems, the initial approach to the
problem was to try to dampen the amplitude of
All of the evidence I have been able to assemble and all of the analyses I have carried out lead vibration at the end of the crank. If we examine
me to conclude that the Allison dampers were not Figures 5 and 6 we see that the 1st order one nodeor the 6th order two node could have induced
necessary to protect the crankshaft, gear box or
vibration at about 2,600 rpm, hence the dampers
propeller drive systems. The only factor that
could explain such a requirement would be inher- for both modes of vibration?
What finally worked was increasing the wall
ently less damping in the Allison engine than in
the Merlin. This, it seems to me, is highly unlikely thickness of the extension shaft. This would have
stiffened the shaft and increased the one node
for the reasons given above.
of its crank. Dixon Smith claims the pendulum
dampers would some times fly off at high engine
speeds and wreck the engine. Consequently someracers removed the dampers. We do know thatwear was a problem with the damper pins andthe holes they operated in. These would gall in
relatively short periods of time due to high contact (Hertz) stresses. The NACA investigated the
problem and reported on it in 1945 ( Reference 7).
They recommended an increase in the pin andhole size and ran tests to show that it worked. It
isn’t clear when, if ever, this fix was introduced in
production engines. Their design was for a takeoff speed of 3,000 rpm. The re-designed pins
reduced the stress from 146,000 psi to 119,100 psi.
At 4,500 rpm the stress would be back to178,600
psi. The effect of wear on the pins is to de-tune
the dampers and enough wear could cause themto break.
I have heard no evidence that the sudden overspeeding phenomenon in hydroplane racing
caused problems with Allison’s supercharger
drive system as it did with the Merlin’s.

Fig 10. AllisonV-1710-E as used in the Bell XP-39

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natural frequency enough to stop it from exciting
the resonance in the cam/magneto drive system.
Note that in Figures 1 and 3 the one node natural
frequency of the V-1710-E was only 44% of theclose-coupled V-1710. The change in the extension
shaft stiffness would have had a much smaller
effect on the two node natural frequency but it's
impossible to judge at this point in time whichmode of vibration was causing the resonance. My
guess is that it was the one node. Changing thenatural frequency would seem to have been theobvious thing to do first, rather than trying to
reduce the amplitude of vibration at the same frequency but we weren't there and shouldn't judge.
Many engineers, myself included, have been insimilar situations where "everything but the
kitchen sink" is thrown at a problem in an attempt
to get the product to the customer on time andperforming as promised.
Once the dampers were designed and installed
they were left in all subsequent versions of theV-1710 and provided an additional benefit. The
torsional excitation of the crankshaft would have
had an equal and opposite reaction on the
crankcase and, therefore, on the airframe in which
it was installed. A pilot with experience of both
the Allison and the Merlin in the same air-frame(if such a person exists) could possibly commenton this.
The fact that Allison was a General Motors
company may have contributed to the adoption ofdampers. By the late thirties many automotiveengines were equipped with dampers and theimpetus had probably as much to do with providing a vibration free vehicle as with protecting the
crankshaft. We know that the General Motors
Research lab was involved in V-1710 development and contributed very significantly to the
optimization of the crankshaft design. It seems
quite likely that they may have encouraged theadoption of dampers. They are simple and do not
add much weight to the engine.

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Acknowledgements
I wish to thank Dan Whitney and Kim
McCutcheon for bringing this topic to my attention and providing encouragement and helpful
information along the way. Dan’s encyclopedic
knowledge of the Allison engine and his collection of otherwise unattainable information madethis effort possible.
Dixon Smith’s generous contribution of time torelate his racing experience to someone who has
never taken a wrench to an aircraft engine ismuch appreciated and I was very gratified to beable to get it in print.
Dave Piggott of the Rolls-Royce Heritage Trust
provided valuable insight into Rolls’ activity inthe area of torsional vibrations.

References
1. Carter, B. C., & Forshaw, J. R. “ TorsiographObservations on a Merlin II Engine, using a SerratedCondenser Pick-Up, with Five Different Pitch Settings of thePropeller Blades.” Royal Aircraft Establishment Reports andMemoranda No. 1983, July, 1943.
2. Hazen, R. M. &Monteith, O. V. “Torsional Vibration of
In-Line Aircraft Engines”, Society of Automotive Engineers
Journal (Transactions). V.43, no.2, Aug., 1938.
3. Torsional Vibration Survey of the Packard Rolls-Royce
V-1650-1 Engine. Army Air Forces Material Center. Report# EXP-M-57-503-632, June 5, 1942.
4. Monteith, O.V., “Detail Design, Analysis andTorsional Vibration Analysis of the Allison Model V-1710-E
Aircraft Engine. Air Corps Type and Model V-1710-17. (Thisreport was prepared for the Army Air Corps on August 15,
1939 to satisfy contract requirements).
5. Whitney, Daniel D. Vee’s for Victory! The Story of theAllison V-1710 Aircraft Engine, 1929-1948 Schiffer
Publications, Atglen, Pa., 1998.
6. Raymond, R.J., “Aircraft Performance Analysis at
Rolls-Royce ca. 1940. March, 2011. Aircraft Engine Historical
Society, www.enginehistory.org.
7. Meyer, Jr., Andre J., “Elimination of Galling of
Pendulum Vibration Dampers Used in Aircraft Engines”.
NACA Wartime Report ARR No. E5G31, August, 1945.

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