A strain gauge torsionmeter for ship shaft systemns

A STRAIN GAUGE TORSION}IETER FOR SHIP SHAFT SYSTEMS

by

H.B. Boyle, C.Eng., A.M.I.Mech.E., A.M.R.I.N.A.

Introduction

The measurement of propulsive power is required at least once in the life of a ship, and quite often the measurement of power or torque is made regularly as part of a general study of performance. This paper describes the experience gained on one type of torsiorimeter, namely that using wire or foil resistance strain gauges.

In 1957, Ship Division of the National Physical Laboratory started upon a research programme involving the need for a torsionmeter to measure torque on small diameter

(k

in) shafts. Because of the small diameter involved, (making most

other systems impracticable) together with the general strain gauge experience of the Division, it was decided to use a strain gauge system. Since that date a number of shaft systems have been gauged, all the while building up a considerable

background knowledge. Critical examination of the results has led the author to believe that a strain gauge torsionmeter is as good as any system for all practical applications and has a number of advantages over most.

Choosing a Torsiônméter System

Ship torsionmeters, almost without exception, use the propeller shaft as the flexural element; the possibility of using a special shaft section, preferably weakened to obtain higher stress concentration or larger deformation is almost non-existent. Even with this limitation many measuring systems have been proposed and quite a number are in regular use. (References 1 and 2).

The initial choice of torsionxneter may depend upon a number of factors such as cost, commercial availability, or convenience of use. Thus when choosing a system it does not necessarily follow that the right system is chosen. For this reason

there must be a critical examination, at regular intervals, to ensure that the best engineering methods are being used.

From simple theoretical considerations there are only two basic systems available if it is assumed that the shaft is to be the transducer element of any torsionmeter system, and some form of mechanical deformation is to be measured. Consider the standard engineering relationship.

T q GQ

Jr

1

this may be rewritten as

2GeJ GQJ T

r 1

where T = Torque

J = Polar moment of inertia

q = Sh.ar stress r = Radius G = Shear modulus Q = Angle of twist 1 = Length

-2-where e = direct strain.

Assiuning J,l and r are known constants, then a chosen system must measure

either:-Method (1) GQ or Method (2) Ge

Therefore from theoretical considerations, each system requires the determination of two unknowns. _{(G is normally measured by ultrasonic methods - Ref 2).}

Besides these theoretical considerations, there are a number of practical problems to be solved. These may be divided into three sections,

namely:-Choice of measuring transducer.

Signal transference from shaft to measuring system. Suitable accurate measurement and read-out.

The electrical measurement system has almost suppressed the use of direct optical or mechanical systems, probably for reasons of convenience, remote read-out and the advantage of electrical magnification being the most important.

Let us then consider the engineering problems for a typical electrical torsionmeter system.

(1) Choice of measuring transducer

Consider the two methods mentioned

above:-Method 1 (GQ) Normafly such a system does not measure the angle Q, but a linear displacement having an electrical signal proportional to tan . In order that

this measured displacement is large, the two points which define 1 should be as widely separated as possible. If 1 is very large it is necessary to measure the displacement from stations mounted on the ships structure. Unfortunately a ship cannot be considered a rigid structure and with such a system large errors are

T qJ GQJ

r

Since shear stress cannot be measured directly, it is necessary to measure the principal strains resulting from the shear stress, hence

'-3

likely to occur due to hull flexing. There is the added problem of shaft

couplings for which the stiffness may be difficult to determine. _{If the} displace-ment is to be measured with reference to two points on the shaft, then 1 will probably be limited to a maximum of ₃ ft. This is due to the conflicting require-ments of producing a light easily portable system with sufficient rigidity not to be influenced by shaft rotation. The whole system would normally be calibrated on a dummy shaft subjected to known angles 0 and then clamped to the ship's shaft. The method of clamping which seems to be preferred is by means of half rings bolted together. Although this system gives rise to doubts as to the effective clamping position, and hence the length 1, it does provide a very rigid method of attachment.

Method (2) (Ge) The torsional shear stress produces principal strains on helices which are at +5° to the shaft axis. It is therefore necessary to attach some form of extensometer or strain measuring device. Since the measurement of strain is along a curved path, the conventional mechanical extensometer is unsuitable and the most obvious choice for such a measurement is the resistance strain gauge Accurate positioning of the gauge is important, but if strain is being measured

the length over which the measurement is made is not important. However the use of the strain gauge involves one further variable quantity, namely gauge factor, which is the ratio of electrical to mechanical strain.

Electrical Signal Transference

The electrical output from a torsionmeter must be transferred from the rotating shaft to a recording station. There are a number of methods available including such sophistication as telemetering systems. Such systems are

unnecessary 'complications for shaft torque measurements when slip rings using silver rings and silver/graphite brushes ar so readily available. Experience has shown that this type of transfer unit is re] iable and does not introduce extraneous signals if used correctly. The quality of signals which may be transmitted is shown in Fig.1. Here the shaft skin strain is of the order ±ij- microstrain

x

o_6

inch per inch strain). The major fluctuations are due to propeller force excitation and only the superimposed small amplitude, higher frequency ripple can be attributed to extraneous signals.

Electrical Measurement

The mechanical. to electrical signal conversion may take a variety of forms, the most common being voltage, frequency or phase dependent. Although the latter two methods may appear to be preferable in that the

signal

amplitude is unimportant, modern electronic technology has made possible the accurate measurement of

14_

The Strain Gauge System The Difficulties

In any strain gauge system whether for a ship torsionmeter, or for any other purpose, there are a number of difficulties to overcome. These difficulties are

cormnon to all gauge installations so it is proposed to deal only with those that require special attention in the construction of a ship torsionxneter. These may be listed as gauge positioning, temperature effects, determination of gauge factor, gauge bonding and sensitivity to other strain directions.

The correct positioning of a gauge depends on two factors, correct 'marking out' on the shaft of the position lines and the bonding of the gauges to these lines. Even with quite simple methods of marking out it will be difficult not to achieve an accuracy of at least 10 and, with care, an accuracy of better than 1/100 is possible. (Yig.2 shows two lines at 10 angle of separation). The resulting error would be .06% for 10 positional misalignment.

Gauge positioning can only be satisfactorily achieved with gauges where the metal grid of the gauge is clearly visible and the gauge has a transparent backing; most foil gauges are admirable for this purpose. The gauges can now be positioned using a portion of the grid to coincide with the markings on the shaft. If gauges of sufficient size are used it is reasonable to get a system which is correctly positioned to approximately ± 1/10°.

Many wire or foil resistance gauges exhibit temperature strain sensitivity; the use of 1 gauges in a torsionmeter helps to reduce this problem but .does not eliminate it. An elegant solution is to use a gauge where the gauge temperature sensitivity matches and cancels the effects due to the strain induced into the gauge by the shaft thermal expansion. Such a system can give excellent

temperature stability. On one such small laboratory installation which was tested a change of approximately 60 deg C induced an apparent strain of only microstrain. This type of commercially produced gauge is again a foil gauge.

Determination of gauge factor is achieved only by measurement. it is the accuracy of this measurement which has the most significant effect on the accuracy of the system. (A simple device, as shown in Fig.3 enables the gauge factor to be derived). At one time it was felt that gauges should be bonded to a test beam at the same time as those bonded to the shaft, since it was thought that the gauge factor might be influenced by environmental conditions. However experience has shown that this is unnecessary.

The modern manufacturer carries out quite exhaustive tests to establish the gauge factor of a batch of gauges, (Ref.3) and as long as the manufacturer's

recommended bonding techniques are foflowed, this gauge factor may be used without recourse to any checking.

5

The arrangement of the gauges to measure torque is theoretically insensitive to the two other main sources of strain, i.e. shaft compression due to thrust and shaft bending due to shalt weight or misalignment of shaft bearings.

In practice the thrust sensitivity is small if good quality matched pairs of

gauges are used. A typical shaft designed to carry a torsional load of 12,000 lbf ft resulting in 250 microstrain gave, when loaded to 10,000 lbf thrust, an output of less than j microstrain.

Although the shaft skin strain due to bending may be large compared to that for thrust, errors are only likely to occur in establishing the torque zero when the shaft is at rest since the bending strain will be cyclic with the shaft rotating and thus the variation of signal may be averaged. (This is usually achieved electrically by using a recording device having a low frequency response indication). Errors in determination of the zero position may be obviated by averaging the readings obtained with the shaft slowly rotating. This technique also reduces any residual torsion in the shaft between bearings.

Discussion of Equipment and Techniques

The four case histories presented here, summarised in Fig.k, have been picked as each high-lights some particular aspect of the overall problem. In all cases the measuring system was identical, and is described below.

1. Neasuring Equipment

This was a strain indicator of commercial make using a carrier system of strain gauge bridge excitation, a 5 volt supply having a frequency of 2,500 or 1,000 cycles per second. The system operates either as a direct reading meter device or as a null balance system. A sensitivity switch changes the value of the full scale deflection of the meter in ranges of 30e, 300te, 1,000i.te and 10,000e steps. The null balance is achieved from two switch positions of 1O,000e and 1,000e steps plus a continuous dial adjustment totalling 2,000.ie marked in 10e intervals. This latter dial can be read to better than ±4te. The strain steps referred to assume one active strain gauge having a gauge factor of 2. The readings for a four-arm torsional strain gauge bridge must be modified as

follows:-Strain = Reading No. of Gauges 2 x Gauge factor

-6

The absolute accuracy claimed by the manufacturer for this measuring system is as follows:

-Accuracy as Null Balance Device ±% Accuracy as Direct Reading Device ±1%

(Taken from Manufacturer's General Catalogue

1966).

Experience with a number of these strain indicators has shown that the order of repeatability and accuracy is certainly higher than quoted. Provision is also made for a rectified signal to be available from two output terminals, voltage being related to the full scale deflection of the meter. Internal calibration control and drift monitoring facilities are also provided. During calibration the null balance system is used, but during measurement it is more convenient to use a combined technique whereby balance is achieved in 100te steps, the residual being read directly from the meter.

2. Slip Rings

Both commercial and "homemade" slip rings have been used successfully, the choice is one of convenience, the commercial ring being easier to fit, the home-made or constructed ring being about a quarter of the cost. Both types of ring mounted on a shaft insert are shown in Fig.5. Wherever these units have been used

they behaved well and have shown no detectable wear. All four cases referred to in Fig.k used. commercial slip rings and in case III a constructed unit on one shalt was used for comparison purposes. In case II, the rings and brushes were

subjected to an estimated 1,000,000 revolutions without any visible deterioration. In cases II and III the same commercial rings were used, the difference in shaft diameter being accommodated by semi-circular packing pieces.

In case III, because of the gauge position it was possible to take zero torque readings both when the shaft was stationary 3nd rotating at 380 r.p.m. The

difference between the zeros, stationary and rotating was not measurable and must therefore have been less than 1 microstrain. This "rotating" zero was recorded during single engine trials, and the trailing propeller caused the disengaged shalt to rotate.

In cases II, III and IV the rings, whilst in use1 were subjected to

unintentional contamination by various fluids. Case II was positioned under a fueld pipe leak, case III having one slip ring assembly splattered with grease from a gear coupling, and in case IV, the drive gear box developed a leak which spilled oil over the slip ring whilst running. In no instance was there a

noticeable effect on the readings, there being no erratic change of signal level nor change of shaft zero from the original clean state.

-7

3.

Brush Gear

In all cases commercial brushes made of silver/graphite were used. These brushes, two per ring, were supported on beryllium copper canti].ever springs. The brushes were attached at the connection end to a common insulating block, arid then angled out to the shaft. The brush contact points were thus at about

600

separation around the shaft. This arrangement made for very easy setting up and removal. Positioning, to obtain the correct spring pressure, was not critical and the unit could be disengaged whilst the shaft was rotating.

k.

Strain Gauges

In case I the gauges used were paper-backed, grid wound iridium platinum wire gauges. These were chosen for the high gauge factor (6). This trial was part of a research into propeller excited vibration (Ref.k). Under these circumstances it was more important to record the fluctuating strains with good signal to noise ratio than to achieve a measurement of mean torque to a high order of accuracy. However, as an AEI torsionmeter, fitted by the British Ship Research Association was

available, it was decided to compare results obtained from both meters. The type of gauge used was very sensitive to temperature change and to reduce this effect careful gauge resistance selection was necessary. Two torque bridges were fitted in the same axial position, but separated by 90° in the radial direction. The slip ring assembly was mounted over the gauge position. The gauges were bonded with rapid setting cement. Due to the extreme cold, local heating was introduced by means of a butane gas burner in the shafting compartment of the ship. Whilst this raised the shaft temperature from 0°C to a more acceptable temperature (about 15°C), the humidity became extremely high, any unprotected steel rusting rapidly. In case II, paper-backed Karma wire grid wound gauges were used. These gauges were bonded on to the shafts out

in

the open alongside the ship's slipway. Electric heaters were installed near the gauge positions and the whole covered with a tarpaulin. During the 2k hour curing time it rained, again giving extremely high humidity conditions. The gauges were then waterproofed and calibrated, and the

shafts installed in the ship. Approximately 2 months elapsed between calibration and use.

In cases III and 1V1 epoxy-backed, etched foil self-temperature-compensated gauges were used and bonded under laboratory conditions.

5.

Gauge Glue

These are listed

in

the case histories but special mention should be made about Araldite Twin Pack. This seems to be an. ideal strain gauge glue for field work. Experiments

in

the laboratory have shown it to be as good as the conventional

-8

epoxy strain gauge resins, and it does not seem to be too critical as to the ratios of the two part mix. Small quantities can be mixed easily and it can be purchased at any local ironmongers. One word of warning; difficulty has been experienced on diameters below 1 in. Under these circumstances the gauged units experience very undesirable hysteresis. It is suspected that this trouble is due to a thick

deposit of glue over the centre section of the gauges If this is so it probably stems from the gauge bonding technique, where little or no bonding pressure is applied.

6.

Revolution Counting

For case I, in conjunction with the slip ring manufacturer, NPL devised a

method of pulse counting. This consisted of a segmented ring composed of alternative silver and epoxy resin segments, the segments being incorporated into the slip ring

construction. The silver segments were joined together electrically during

manufacture and earthed to the shaft during installation. A capacitance probe was then used as a pulse generator, the pulses being fed to an electronic counter via a suitable amplifier system.

The system has proved to be very reliable. It has a number of advantages over reflective light, photoelectric or inductive systems, since it is not affected by oil or dirt and the pulse size is not related to the shaft rotational speed. In

case II, NPL were not required to measure power, only torque, and so no tachometer was fittede

In case III, the high speed diesel engines were fitted with shaft pick-up points suitable for hand-held mechanical tachometers. For this particular trial, where the engines were speed governed, a hand-held tachometer was considered more

convenient and was used.

In case IV, the engines were supplied 'ith electronic counters.

Measurements and Results Case I

Readings were taken of the outputs fron the two strain gauge bridges, the AEI torsionmeter and the NPL shaft tachometer. These readings were taken over the major part of the shaft revolution range, whilst the ship was proceeding along the English Channel. The weater conditions were poor and this was reflected in the general random scatter of results, probably due to ship motion which resulted in

continually changing shaft rotational speed. This was aggravated by aU readings being taken in serial form; due to limitations of equipment then available it was necessary to switch each strain gauge bridge, in turn, into a single strain

indicator. Since records of fluctuating strain were being taken there was, of necessity, a considerable time lapse between measurements of the two bridges.

9

It must be stated that the experiment was not performed essentially to compare torsionmeter accuracies but was part of an overall vibration study. The results obtained from the strain gauge and AEI torsionmeters are shown compared in

torque Fig.6. They have been plotted as

(shaft rotational speed) versus shaft revolutions; the latter plotting provides an expanded vertical axis. The AEI torsionmeter readings were converted into torque on the basis of a formula

provided by BSRA. The formula involved the use of a correction factor obtained by calibration. This calibration is shown in Fig.7.

Case II

The shafts for this ship were calibrated by means of a very simple calibration rig; the results of the calibration are shown in Fig.8.

To calibrate, each shaft in turn was laid on a concrete floor and restrained at each end in the vertical direction by two metal straps grouted to the floor. Single torque arms were bolted to each end of the shaft, through the normal shaft

coupling flanges. One arm was bolted to the floor, the other arm being attached at its outer end to a tripod via a tension screw jack and a measuring hydrostatic

capsule (accuracy

±5

of full scale deflection). The torque applied was equivalent to the capsule reading times the torque arm distance of 10 ft. To reduce the

errors due to friction, the rig was hit at the metal strap positions with a heavy mallet at each load condition.

Case III

A simple calibration rig was constructed at NPL and the constructional details are shown in Fig.9. The load measuring capsule was similar to that used in

Case II. The torsionmeter units were calibrated twice and used for ship trials on three occasions over a period of eleven months. Note was taken of zero shift and overall repeatability and these are discussed later. The calibration was made at Ship Division, Feltham, whilst all trials work was carried out on the Solent. Calibration results are shownin Fig.10.

Although no other torsionmeter was available, the engine manufacturer had calibrated the engines on a brake at the factory, arid, during the trial, records of r.p.m. and various engine parameters were taken. From this information

horsepower figures were derived. A comparison of manufacturer's and NPL values of power are shown in Fig.11.

Case IV

In this series of trials the shipbuilder and NPL produced 12 shaft torsion-meters, each shaft being fitted with two strain gauge bridges. The shafts were calibrated in both the port and. starboard turning directions in the rig mentioned

10

-for Case III. Three shafts selected at random, as far as the strain gauge bridges were concerned, were then sent to the engine manufacturer to take part in engine and gear box trials. This enabled a comparison to be made between a calibration of a simple static type and results obtained froi a Heenan and Froude water brake under running conditions. The results of these calibrations and trials are shown in Figs.12 and 13.

Discussion of Results Case I

The results, Fig.6, show good agreement between the two measuring systems. It was to be expected that the two strain gauge bridges would be in better agreement with each other than with the AEI torsionmeter, since they were measured at closer intervals of time. However, there is no doubt that the best smooth line, as

defined by the strain gauge system, was confirmed by BSRA readings. The results of the AEI torsionmeter static calibration are shown in Fig.7, and it is interesting to note the change of calibration factor with respect to output reading; this variation is approximately 3%. Although the readings were converted to torque by use of this correction curve and calibration formula, it is preferable to have a torsionnieter which has a more linear relationship between torque and output reading.

As will be seen in Cases II, III and IV, strain gauges meet this requirement.

Case II

The object of including the results of this trial is to show the possibilities of calibration of smaller shafts by simple means. The calibrations in Fig.8 show the repeatability that is possible on two separate shaft systems.

The general shape of the calibration curve is influenced by a number of factors; friction within the calibration system, load capsule calibration errors, and the inaccuracies caused by the displacement of the torque arm through a large angle under load. This displacement was caused by the twisting of the shaft under torsional loading and the flexibility of the torque arm. Since the jacking anchor

point

was fixed, there was an effective change of length of the torque arm. (See

Fig.lk).

The friction losses almost certainly produce the steeply falling portion of the curve in the lower part of the calibration, whereas the rising portion of the curve is due mx4inly to the angular displacement error mentioned above.

Case III

This set of results (Figs.10 and ii) is excellent, both the repeatability achieved over a period of months and the stability of the zero indicate that the

strain gauge system offers long term stability for torsionmeter systems. The zero scatter was of the order of ±1 i/'+% of Full Scale Reading spread over a period of 8 months. This period involved two trials on separate ships, two calibrations in the laboratory, and the use of two similar strain indicators. Using the shaft diameters and gauge factors, the calculated values of modulus of rigidity of the two shafts are 11.91 x 1O6 lbf/inch2 and 11.97 x lbf/inch2, which are

satisfactory values for steel, lying within the spread of modulus measurement made by the BSRA (Reference 3). Again, there is the steeply falling curve at the lower end of the calibration due to friction, but even accepting this possible source of error the dynamometer has a linearity of better than ±1% over half of the torque range and better than ±2% over four fifths of the range. It is now normal

practice to ignore the lower portion of the calibration curve and to establish the calibration factor by extrapolation of the upper part of the calibration towards zero. In this particular case this would give calibration factors of approximately

6.25 and 6.28 lbf ft/division. The correctness of such a decision was confirmed in Case IV when the calibration curve of similar shape was treated in the same way. The results are discussed below.

Case IV

During the gauging and calibration o± these shaft inserts, the opportunity was taken to determine the effect of the proximity of the flange upon the gauge output. These particular shaft inserts were 8 inches between flanges and five strain gauge bridges were bonded to one shaft, the gauge positioning being at equal intervals along the shaft The shaft was loaded torsionally and. each bridge output recorded. Only the bridge nearest to the flange gave a significantly different output of the order of 2%. The others gave readings which were within the tolerance of the gauge factors for this particular set of gaiges.

The twelve shafts were calibrated

in

the rig described for Case III and then three of the shafts were used during engine trials on a test bed.. The calibration results for these three shafts and the comparison of brake and torsionmeter

readings are shown in Fig.12 and tabulated in Fig.13.

The shaft static calibration shows the now expected shape for this type of calibration, and it was decided to extrapolate the best fit straight line obtained from the higher portion of the calibration, to derive the lower portion. This calibration line was used to derive the torque values for the brake/torsionmeter tests. In this comparison only four torque values were more than 1% in disagreement with the brake results.

The biggest difficulty with this type of comparison is the difference in response of the two systems. The brake recording dial is deliberately damped to give a stea&y reading and this damping produces a significant delay in recording

12

-changes in torque. Since the strain indicator is read at the same time as the brake dial, errors must occur, This is the probable reason for the apparent scatter of results even at apparently constant brake torque conditions.

Conclusions

Methods of measuring ship shaft torque have always been a controversal subject, each method having its own champions. However a number of systems produce equally accurate results when installed by skilled technicians. The general accuracy of all the systems is in the order of ±2% of maximum torque and under favourable conditions ±1% is possible. This is probably the limit if the ship shaft is used as the spring element. Greater accuracies could be achieved by special calibrated shaft inserts, but the production of such units for large

torque values would be costly.

The advantages of the strain gauge system lie in other directions and the more important can be listed as

follows:-Shaft size

Many torsionmeter systems are not suitable for attachment to shafts of small diameter. The strain gauge does not suffer from this limitation. Slip rings of the commercial type can be made or adapted to suit a specific shaft size, but the temporary ring constiuction is limited to a minimum diameter of approximately 3 inches unless pre-formed during manufacture.

Installation time

At first sight it would appear that a "clamp on" torsionnieter would be quicker to install than a strain gauge system. However in the author's experience this is not the case. On one occasion the author and one assistant fitted two complete

strain gauge torsionmeter systems in less time than it took to install a "clamp on" torsionmeter fitted to the same shaft.

If slow drying glues are used the installation time will take 2k hours although the actual time "on site" is

only

about

6-8

hours. üth quick drying glues or impact adhesives the total "on site" is, obviously,

6-8

hours, These times assume that all the necessary facilities are available.

Skilled labour

All torsionmeter systems require skilled labour to ensure a proper

installation. Strain gauges are used on an amazing variety of "field" experiments and because of this there is probably a greater labour pooi capable of fitting a strain gauge torsionmeter than all other torsionmeter systems put together.

13

-+. Cost

The price of a torsionmeter installation will depend upon the sophistication of the system used. However, in its simplest form, the strain gauge torsionmeter is cheaper than any other system.

Such a system would employ D.C. voltage bridge excitation with the output measured using a D.C. microvolt meter calibrated in iv/v or direct strain. With temporary slip rings such a system would have a total cost of less than 1OO plus installation costs. Such a system was adopted by the White Fish Authority and described by them in Reference 5.

Rotational speed

The gauges, having negligible mass, are neither affected by high rotational speed nor influence the shaft. The shaft speed is limited only by the behaviour of the slip ring assembly. Since the ring mass is small and balanced about the

shaft axis, the ring speed is limited only by the rubbing speed between ring and. brush. Modern techniques are such that "noise free" running is possible at speeds far in excess of any likely ship shaft speed.

Reliability

It is likely that there have been more experiments abandoned due to strain gauge system failures than any other type of electrical strain or displacement measuring device. Because of this the strain gauge has received much bad

publicity. This is unfortunate since the majority of failures have occurred when no other system could have been used to perform the required task. These failures

are usually caused by installation damage, mainly wiring, and it is this type of failure which has caused much of the criticism of strain gauge systems. In fact the author cannot remember an occasion where failure could. be directly attributed to gauge failure unless due to damage from some external source.

A strain gauge torsionmeter suitably protected

rtfl

give a high degree of reliability, with the additional advantage that a second gauge bridge can be

bonded to the shaft to be available in the event of damage. This extra strain gauge bridge takes little extra time to install and one bridge can be used as a measuring check on the other to confirm satisfactory installation of both systems.

The strain gauge system, like all others, has a number of disadvantages, and for satisfactory operation the effects of these disadvantages must be reduced to a

minimum:-(a) Damage during installation

A strain gauge is a fragile device and is easily damaged during the

installation period. Every care must be taken to ensure adequate protection at all times.

1k

-Earth resistance

The strain gauge bridge will cease to measure accurately or even have a linear calibration if the bridge earth resistance is allowed to fall below 100 megohms. This can lead to difficulty during installation unless adequate precautions are taken.

Brush gear damage

The recommended use of cantilever springs for the brush gear requires that adequate protection be given to them since they are fragile.

Machining accuracy of the shaft

In the case of shafts which are oval or for hollow shafts where the bore is not concentric, errors occur

in

trying to establish the values of torque from the strain data. Under these circmtances calibration is the only suitable solution in obtaining accurate results.

Cable resistance

If long runs of cable are used due allowance must be made for the voltage drop along the bridge input supplies. This was necessary in Case IV where the recording station was situated on the bridge of the ship. This is done by

comparing the gauge bridge resistance to cable resistance and correcting the strain reading proportionately.

Acknowledgement

The work described above has been carried out at the National Physical Laboratory.

The author wishes to acknowledge the co-operation and assistance of the following organisations

-Bristol Siddeley Engines Ltd. British Ship Research Asfociation. D. Napier and Sons Ltd.

National Institute of Oceanography. Vosper Ltd.

The author is also indebted for the help of staff of Ship Division, NPL; in particular, to members of the Equipment Group.

-

15

-References

(i) COOK, R. 1951 Marine torsiozimeters and thrust meters. Trans. I. Mar. E.,

Vol.63, p,ll5.

MORRISON, T.

1966

Recent developments in the measurement of propeller shaft torque and thrust.

Triu. I. Mar. L,

_{Vol.78, p.193.}

COIJTTS, J.A.

₁₉₆₇

Manufacturers methods of calibrating bonded resistance strain gauges. Journal of the British Society of Strain Measurement,

Vol.3,

No.2.

(1,) _{SIL1TERL]AF, A.., MARWOOD, W.J., BOYLE, H.B.}

_196k

_{Some ship and model}

measurements of unsteady propeller forces. Trans. R.I.N.A.,

Vol.106,

p.klS.

(5)

BENNETT, R. and HATFIELD, M.

1966

Development of strain gauge techniques for the measurement of propeller shaft torque in distant water trawlers. The Journal of Strain Analysis, Vol.1, No.2.

FIG. 3.

ALL

FIGURES

QUOTED

ARE NOMINAL

VALUES

*

DIPFERENY

INSTRUMENTS

USED DURING

INITIAL

CAL.

AND

TEST

BED TRIALS

FIG. 4. SUMMARY

OF

CASE

HISTORIES

CASE NUMBER

1 2 3

4

TYPE OF SHIP

OCEANOGRAPHICRESEARCH VESSEL HIGH SPEED ROUND BILGE

CRAFT

HIGH SPEED

liOft. MOTOR

BOAT

HIGH SPEED

PLANING

CRAFT

Z

0

4

o

o

w

a.

I-LI.

4

U)

SHAFT DIA.

9 InS 51/4 InS

49x 40(HOL) 51/4X 31/4(HOL)

MATERIAL

STEEL. STEEL STEEL STEEL

MAX. R.P.M.

190

42.0

1,000

1,700

MODULUS OF

RIGIDITY

(lbf/1n2)

11B6x106

UNKNOWN

UNKNOWN

UNKNOWN

MAX. LINEAR

STRAIN

-AU GE

220/..Le

530ue

570,u.e

25o,e

MAX. SHEAR

STRESS

(Ibf/Ifl2

_'

5,300

12,700

13,700

6,000

N OF SHAFTS 1 2

2

=

3 _j

4

I-a

z

0

4

-J

z

-GAUGE TYPE PAPER

IRIDIUM

PLATINUM WIRE PAPER

KARMA

WIRE EPDXY

FOIL

EPDX.Y

FOIL

GAUGE FACTOR 6 2

2

a

N°' MEASG.

BRIDGES

2

1.- PERSHAFT a- PER SHAFT 2-PER SHAFT

BRIDGE NER.GISATI0N

1,000 C/s AT

5-VOLTS

1,000 c/s AT

5-VOLTS

2,500 c/S AT

S-VOLTS

1,000 AND

*

_{2,500 cls}

AT 5-VOLTS

GAUGE GLUE PHILLIPS

STRAIN GAUE

CEMENT

ARALDITE

TWIN PACK

ARALDITE

753

ARALDITE

753

SLIP RING

TYPE COMMERCIAL SPLIT COMMERCIAL SPLIT COMMERCIAL SPLIT E4 CONSTD

COMMERCIAL

SPLIT

METHOD OF

CALIBRATION

OR COMPARI SON OF RE SULTS

A.E.1.

TORSIONMETER FITTED BY B.S.R.A. SIMPLE

CALIBRATI3N

RIG CONSTD

'ON SITE'

CA LI B RATION

RIG, (FIa9.)

CALIBRATION

(FIG.9.) PLUS

CHECK ON TEST BED

SPECIAL

0 M ME NIS OUTPUT

MEASD. BY

NULL BALANCE NONE

SPECIAL

SHAFT

INSERTS

SPECIAL SHAFT

INSERTS

8,500

8,000

('I

7,500

Ui

±10

7,000

Lii

0-i

0>

6,500

6,000

0

0

X

STRAIN GAUGE BRIDGE. 1.

STRAIN

AUE BRIDGE. 2.

o

A.E.1.

TORSIONMETER

10

ao

SHAFT

REVOLUTIONS/ SEC (n)

30

FIG. 6. COMPARISON

OF

STRAIN

GAUGES AND

A.E.1.

TORSIONMETER

z

0

lii

1O5

uI<

'czW

a:

_I04

tow

ow

103

I-U,

jg

wo

0

0

I-101

100

30

40

50

60

70

80

90

100

110

TORSIONMETER

READING (DIVISIONS)

FIG

7.

CORRECTION

FACTOR

USED

FOR

A.E.1. TORSIONMETER

1700

1650

1.00

1550

1500

X

SHAFT. 1.

-I-

SHAFT. a.

200

400

600

800

1,000

1,200

1,400

READING (DivisioNs)

FIG,

_8.

CALIBRATION

OF

SHAFTS

CASE

HISTORY

II

////////////////////////////////// ///

LIFTING JACK.

- HYDROSTATIC

CAPSULE.

LOADING GUIDES.

FIG 9.8. END ELEVATION

z

2 660

>

a

4-9-

650

"S

640

0

I-0

LL

670

630

620

0

x

El

+

0

x

+

_SHAFT.a.

9

0

1g

x

0

_SHAFT.1.

O

SHAFT ()

CAL.1.

El

SI-IAFT ()

CAL.a.

ELAPSED TIME BETWEEN

X

_{SHAFT ®}

CAL.1.

CALIBRATION .1 .

E CALIBRATION.2.

+ SHAFT ®

CAL. a.

APPROXIMATLY 2 MONTHS

2.00

400

600

800

1,000

1,200

1,400

1,600

1,800

2.,000

BRIDGE

OUTPUT

(DIVISIONS)

FIG. 10.

SHAFT

CALIBRATIONS

RESULTS ARE FOR THE TOTAL

POWER FROM BOTH ENGINES

1,000

2,000

3,000

4003

5,000

6,000

N.P.L. POWER (HP)

FIG. 11.

COMPARISON OF tLP.L.

AND ENGINE

MANUFACTURERS POWER MEASUREMENTS

+

x

0

0

0

+

BRIDGE

o

-100

200

300

400

500

600

700

BRIDGE OUTPUT

(DIVISIONS)

FIG. ia. CALIBRATION

OF

SHAFTS

CASE

HISTORY

(THESE

SHAFTS USED

FOR ENGINE TEST BED TRIAL)

600

--0-w D

0

w

J

0 a-

U-0

0'

'-4 .1-I

1250

1240

1230

ta20

0

I-tL

z

1210

0

1200

-J 4-)

iiqo

SHAFT a

BRIDGE A. SHAFT 3

B.

SHAFT 3

BRIDGE

B.

SHAFT 12

BRIDGE

900

10O

FIG. 13

RESULTS OF ENGINE TEST BED TRIAL

ON THREE SHAFTS

CASE IV

R.P.M.

TORQUE

lbf ft

TORQUE

lbf ft

(R

TORQUE

x ioo)%

BRAKE TOR W SHAFT 3 BRIDGE A 695 1826 1846 +1.07 715 362'-F 3606 -0.1+8 752 5+9O 51+1+3 -0.86 752 51+65 51+30 -0.66 751 5610 5581+

-o.k6

900

721+3 719k -0.62 SHAFT 3 BRIDGE B 1050 8226 8205

-0.25

1050 8262

₈₂₂₅

-0.1+1+

1202

9221

9132

-0.98

1350 10355 101+39 +0.81 1350 10211 10288

_0.75

1520 9996 ioiok +1.02 171+3 12323 12371 +0.39 171+3 121+51+ 12502 +0.38 171+3 12570 121+37 0.51+ 171+3 12311 12371 0.1+8 17+3 12358 121+10 0.1+2 171+3 12299 121+10 +0.89 171+3 12335 12397 +0.50 171+3 12323 12381+ +0.1+9 171+3 12323 12401+ +0.65 1525 12335 12239 -0.78 1525 12323 12338 +0.12 SHAFT 12 BRIDGE B 1352 10120 9993 -1.20 1525

10022

10141+ +1.21 1525 1191+3 12010 +0.51+ 1683 12099 12187 +0.72 171+3 12321 12371 +0.59 SHAFT 2 BRIDGE A 1350

io86

10656

+0.65 1525 10112 10190 +0.76 1525 12001 12111 +0.91

1683

12001 12111 +0.91 171+3 12298 1231+5 +0.38

ZERO FORCE.'

POSITION

z

A

0

SHAFT

AXIS

THE TOR.QUE

ARM AO, WHEN LOADED BY A

FORCE F

ROTATES TO A NEW POSITION

A10

DUE TO TWISTING OF THE SHAFT. SINCE THE

JACKING

AXIS

IS

FIXED, THE EFFECTIVE

TORQUE

ARM IS

ie.

APPARENT TORQUE

= F x

ACTUAL TORQUE

= F x

e1

F x C COS O(

(SMALL

ERROR DUE TO TORQUE ARM FLEXING

IS

NOT SHOWN HERE

IG.14.

EFFECTIVE

CHANGE

OF LENGTH OF