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942478

Acceleration Test Method for a HighPerformance Two-Stroke Racing Engine

Robert J. Kee and Gordon P. BlairThe Queen’s University of Belfast

Reprinted from: 1994 Motor Sports Engineering Conference ProceedingsVolume II: Engines and Drivetrains

(P-288)

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41

942478Acceleration Test Method for a High

Performance Two-Stroke Racing EngineRobert J. Kee and Gordon P. Blair

The Queen’s University of Belfast

ABSTRACT

This paper describes an inertial dynamometersystem which has been applied to the testing of smalltwo-stroke kart racing engines. The dynamometerincorporates a flywheel of appropriate moment ofinertia to simulate the mass of a kart and driver. Thetest procedure involves measurement of the flywheelspeed during an acceleration phase resulting fromopening the throttle. Calculation of the instantaneousflywheel acceleration corresponding to each enginespeed directly gives a measure of the torque andpower characteristics. Performance results, includingexhaust pressure traces, are presented from a seriesof tests conducted on a 100 cm3 kart engine. Theresults are compared with corresponding steadystate measurements recorded on an eddy currentdynamometer. In addition, the measured results arecompared with predictions from a computersimulation.

INTRODUCTION

The outstanding features of the crankcasecompression two-stroke cycle engine are its highspecific power, low weight and mechanical simplicity.Thus, this engine type is well suited to racingapplications in vehicles such as motorcycles, karts,snowmobiles and powerboats.

Development and tuning of two-stroke racingengines typically involves a combination ofconventional steady state dynamometer testing inthe laboratory and timed trials on a racing circuit.However, under typical racing conditions, the engineseldom operates in a steady state condition andengine acceleration rates of up to 7000 rpm/see areencountered. Since the performance of two-strokeracing engines may be significantly influenced by theacceleration rate of the engine, the value of steadystate dynamometer testing is often limited. Hence,greater emphasis is often placed on the timed trialsconducted under dynamic racing conditions.

A number of well known difficulties are alsoencountered during the steady state dynamometertesting of high performance two-stroke engines. Thefirst difficulty is piston seizures due to excessivethermal loading on the piston. Most highperformance two-stroke engines operate close to thedesign limit under racing conditions, where the time-averaged power output from the engine issignificantly less than the maximum available fromthe engine. Consequently, during full load testing, thelimits of the engine may be exceeded during the 20to 30 seconds of running required to record a stablemeasurement. In order to reduce the likelihood ofdamage to the engine, the fuel/air mixture is oftenrichened and the ignition timing is retarded. Theseprecautions reduce the thermal loading, but give anunrepresentative measurement of engineperformance.

A second difficulty arises due to the dependenceof power output on the temperature of the engine andits exhaust system. In the two-stroke engine, thecharging of the cylinder, combustion of the mixtureand the discharging of exhaust gas areaccomplished within one revolution. Consequently,the charging and discharging events overlap andhence the gas flow during these phases is highlydependent on the pressure wave action in the inletand exhaust tracts. Since the velocities of the wavesare in turn dependent on the gas temperatures, itfollows that the output of each engine cycle is highlydependent on the temperature characteristics of thegas resident in the exhaust and inlet tracts. Inaddition, the density of the charge transferred fromthe crankcase into the cylinder is dependent on thetemperature of the inlet tract and the crankcasesurfaces. Consequently, this temperature effect,particularly in air-cooled engines, contributes to poorrepeatability during steady state testing.Furthermore, during a transient acceleration phase,the instantaneous temperatures corresponding toeach engine speed differ from those recorded understeady state conditions. Thus the power developedunder steady state

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conditions may differ from that produced on the racetrack. Figure 1 illustrates the influence of cylindertemperature on engine performance.

Fig. 1 Influence of Cylinder Temperature

A further difficulty arises when the engine is fittedwith a mechanical exhaust timing control device, asthe response of the mechanism to a change inengine speed will incorporate a finite lag time (1). Inorder to maximise performance during acceleration,advanced management systems incorporate facilitiesin which the ignition timing, fuelling and exhausttiming strategies are dependent on the accelerationrate of the engine. For such engines, it isimmediately apparent that steady state testing willnot provide representative performancecharacteristics (2).

Finally, steady state testing usually involvesmeasurement of performance at discrete speedintervals of 500 or 1000 rev/min. Occasionally, asignificant decrease or ‘hole’ in the torquecharacteristic can fall between two measurementspeeds.

From the foregoing it is clear that, in addition tothe practical difficulties, the optimization ofperformance under steady state conditions is unlikelyto give the optimum characteristics for the dynamicracing situation. Laboratory optimization of highperformance two-stroke racing engines ideallyrequires the use of a technique which reproduces thedynamic conditions of the racing circuit. This paperdescribes the development and application of asimple and effective dynamic test method whichemploys the use of an inertial dynamometer.

KART ENGINE

In the European 100 cm3 kart racing classes, thecrankshaft of the engine is directly coupled to therear axle by a chain. Consequently the drive ratio isfixed. The rules do not exclude the use of centrifugalclutchs, but in practice they are only used by somejunior classes. The typical racing circuit is 500 to1000 meters in length, and usually incorporates onemajor straight. Since only one drive ratio is available,the engine is required to operate over a wide rangeof speeds. Typically, the current engines reachspeeds above 18000 rev/min at the end of thestraight and approximately 6000 rev/min at theslowest corner (3).

The results presented in this paper relate to areed valve induction engine which is manufacturedby the lame company and marketed under the Siriobrand. The specification of this engine is given inTable 1 and the engine is illustrated in Fig. 2.

Table 1 Engine Specification

Fig. 2 100 cm3 Kart Engine

INERTIAL DYNAMOMETER

The inertial dynamometer is an alternative testingsystem which measures power under accelerationconditions. This method has many attractions, butthe transient nature of the test procedure requiresthe use of a computer and a data acquisition system.The current availability of low cost computerhardware has eliminated this potential drawback.

At The Queen’s University of Belfast, an inertialdynamometer has been designed by students as theproject element of their engineering degree course.The dynamometer incorporates a flywheel ofappropriate moment of inertia to simulate the mass ofthe kart and

Engine SirioBore 50.0 mmStroke 50.0 mmSwept Volume 100 cm3

Induction Reed ValveCooling AirExhaust Port Opens 92.5° atdcTransfer Ports Open 116.5° atdcTrapped Compression Ratio 10:1Carburettor Venturi Diameter 19.8 mmCarburettor Type Butterfly

43

driver. The test procedure involves measurement ofthe flywheel speed during an acceleration phaseresulting from opening the engine throttle.Calculation of the instantaneous flywheelacceleration corresponding to each engine speeddirectly gives a measure of the torque characteristic.A computer data acquisition system is used to recordthe flywheel speed and to calculate the performancecharacteristics.

DYNAMOMETER DESIGN - The dynamometerincorporates a flywheel which is directly coupled tothe engine by a chain as shown in Fig. 3. The chaindrives the flywheel via a flexible rubber coupling anda shear pin. The function of the flexible coupling is tointroduce compliance into the drive system andthereby reduce the loading on the engine crankshaftdue to torque and speed fluctuations. The shear pinis included to reduce damage to the engine in theevent of an engine seizure. When the pin breaks, theflywheel will continue to rotate on lubricated bushesand thus the stored energy will not be dissipatedthrough bending and breakage of enginecomponents. A disc brake is incorporated in thedesign to reduce the flywheel and engine speedsafter each test. An automotive clutch allows theflywheel to be accelerated prior to engaging theclutch and starting the engine.

An optical shaft encoder is coupled to thedynamometer flywheel to provide a frequency signalwhich is converted to an analogue voltage by anintegrated circuit. This voltage is proportional toflywheel speed and is connected to one input of acustom designed data acquisition system. A secondinput channel is used

to record the temperature of either the crankcase orthe spark plug seat.

The test procedure involves the following:

1. Set engine speed limits for data acquisition(usually 7000 and 16000 rev/min).

2. Start engine.

3. Operate at low speed until the enginetemperature measured at the plug seat reaches75 °C.

4. Switch computer data acquisition system to‘acquire’ mode.

5. Open throttle fully and record flywheel speedduring the ensuing acceleration.

6. When the maximum engine speed is reached,close throttle.

7. Apply the brake to stop the flywheel.

Analysis functions within the computerprogramme calculate the engine torque, bmep andpower. Figure 5 shows a typical speed characteristic.

Fig. 3 Inertial Dynamometer

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Fig. 4 Shaft Assembly

Fig. 5 Typical Engine Speed Characteristic

The torque developed by the engine is calculatedby the following equation.

(1)

where

ωe = Engine speed (rad/sec)

If = Flywheel inertia

Rf = Ratio of flywheel speed to engine speed

Tb =Torque due to dynamometer bearing friction

The torque due to friction in the dynamometerbearings was calculated from a run-down test. Thevalue was found to be in the range 0.13 to 0.15 Nm.From the torque characteristic, the bmep and powercan be easily calculated.

Fig. 6 Typical Torque Characteristic

FLYWHEEL INERTIA REQUIREMENT - Thedynamometer simulates the loading on a kart racingengine during a typical acceleration period. In a kart,the crankshaft of the engine is directly coupled to therear axle and thus only one gear ratio is available.This ratio (Rk) is defined as the ratio of axle speed toengine speed. During acceleration of the kart due toa torque (Te) applied by the engine, the rate ofacceleration is dependent on the mass of the kart(mk), the gear ratio (Rk) and the radius of the rearwheels (rw). Neglecting wind resistance and theinertia of the rotating components, the force toaccelerate the kart is given by:

(2)

Te If= ω· eRf2

Tb+

F mk ωe· Rkrw=

45

Fig. 6 Kart Drive System

The torque applied by the engine is given by:

(3)

Acceleration of the dynamometer flywheel due toan engine torque (Te) is dependent on the flywheelinertia (If) and the ratio of flywheel speed to enginespeed (Rf). Thus,

Tf is the torque applied to the flywheel, and ωf is theflywheel speed. For equivalent loading,

The analysis given above has shown that theMoment of Inertia (If) of the dynamometer flywheel isdependent on the mass of the kart (mk), the radius ofthe rear wheels (rw), and the gear ratios Rk and Rf.The variables mk and rw are determined by the kartspecification and are therefore independentvariables. Rk is dependent on the settings for eachcircuit - an intermediate value 0. 125 was chosen.The design variables are If and Rf.

As illustrated in Fig. 7, it can be seen that aninfinite range of values for If and Rf will satisfy Eq. 5.Consideration of stresses in the flywheel and clutchassemblies aided in the selection of appropriatevalues for If and Rf. - 0.81 kgm2 and 0.25respectively.

Fig. 7 Relationship between If and Rf.

ENGINE TESTING

The inertial dynamometer has been used toevaluate the performance of the 100 cm3 kartengine. Two different exhaust systems were used toillustrate the influence of this component on theperformance characteristics. The exhausts areillustrated in Fig. 8 and had identical overalldimensions. The first was a ‘stock’ exhaust andincorporated a perforated internal convergent cone.The perforations in this cone reduce the amplitude ofthe reflected pressure wave, and are intended to givea broader torque characteristic. The second exhausthad no perforations and has been designated ‘solid’.

Fig. 8 Exhaust System

Comparisons of the power and bmepcharacteristics are shown in Figs. 9 and 10. Theseresults show that the perforated exhaust gave lowerpeak values, but a much broader torquecharacteristic. Examination of the test data showsthat the times for acceleration from 6250 to 15500rev/min were 4.49 and 4.72 seconds for theperforated and solid systems respectively. Thus thebroader characteristic of the perforated exhaustgives a time advantage of 0.23 seconds.

Excellent repeatability was noted throughout allthe tests conducted on the inertial dynamometer -power and bmep measurements were repeatable towithin ± 2%. It is believed that this is due to thestandardised test procedure and the smalltemperature variations which occur during the brieftest. Prior to recording results, the engine waswarmed up to give a crankcase temperature of 35°Cand a temperature of 75°C at the plug seat. Duringthe acceleration test, it was found that thetemperatures varied by less than 2°C and 15°C atthe crankcase and plug seat respectively.

Te mk ωe· R

2k r

2w=

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Fig. 9 Comparison of Power Characteristics

Fig. 10 Comparison of bmep Characteristics

EXHAUST PRESSURE MEASUREMENTS - Themeasurement of exhaust pressure characteristics isan essential element in the optimization of any highperformance two-stroke engine. Usually, thesemeasurements are recorded during steady stateoperation and are fraught with the difficultiesdiscussed earlier in this paper. To avoid thesedifficulties, an alternative method has beendeveloped which allows pressure traces to berecorded during an acceleration power test on theinertial dynamometer.

Since the system was required to record thecyclic variation of pressure at engine speeds up to16000 rev/min, relatively high sampling rates wererequired - at 16000 rev/min, a sampling rate of 96kHzis necessary to give one reading of the pressure andtdc signals for every two degrees of crankshaftrotation. Therefore due to memory limitations, it wasnot possible to sample continuously at this rateduring an acceleration from 6000 to 16000 rev/min.To overcome this limitation, a Computer Boards DAS330 data acquisition system was configured tosample data at discrete speed intervals. An analoguevoltage which was proportional to engine speed was

connected to the first input channel of the board anda customized programme was written to controlsampling of the other input channels. When theanalogue voltage at the first channel reach a valuecorresponding to a required speed, data wasrecorded at the other input channels for apredetermined number of cycles - usually 4. Exhaustpressure data was recorded at 1000 rev/min intervalsfor speeds from 8000 to 16000 rev/min.

A Kistler type 4045 A5 water cooledpiezoresistive transducer was connected via anamplifier and low-pass filter to the second inputchannel. The filter used was a 4th order low-passButterworth type and was tuned to give a cut-offfrequency of 2.5 kHz. This filter introduced a groupdelay of approximately 0.16ms which gave acorresponding phase shift in the pressuremeasurements. For low speed measurements, thisdelay corresponds to a small phase shift - 2.9° at3000 rev/min. However, at 16000 rev/min, the delaycorresponds to approximately 15°. In order to reliablyevaluate the phase shift, the unfiltered signal wasrecorded at the third input channel. Finally, a tdcreference signal from a photodiode was connected tothe fourth channel. An overall sampling rate of250kHz was used to give a sampling rate of 62.5kHzfor each channel.

Fig. 11 Comparison of Exhaust Pressures at8000 rev/min

RESULTS - The results presented in Fig. 11 show acomparison of the exhaust pressure traces recordedat 8000 rev/min. The annotations EO, TO, TC andEC denote exhaust open, transfer open, transferclose and exhaust close respectively. It should benoted that the pressure transducer was located100mm from the exhaust port, thus giving theobserved delay between opening of the exhaust portand arrival of the blowdown pulse.

For both exhausts, the reflected plugging pulse isreturning to the exhaust port too early to give aneffective increase in cylinder pressure at exhaustclosure. The early arrival of this pulse also reducesthe beneficial influence of the suction pulse whichresults from reflection of the blowdown pulse fromthe divergent cone. The resultant effect of therelatively poor trapping pressure and delivery ratio isa torque value well below that produced at the

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optimum speed for the exhaust system, Comparisonof the two characteristics shows that the pluggingpulse reflected from the solid cone had a greateramplitude and was followed by a suction pulse, justprior to exhaust closure. These differences resultedin the 1 bar decrease in bmep noted at 8000 rev/min.

Figures 12 and 13 show exhaust pressure tracesrecorded at 10000 and 16000 rev/min with the solidcone fitted. These traces illustrate the exhaustpressure wave characteristics at the tuned speedand at a speed well beyond the peak torque point. At10000 rev/min, a strong suction pulse is followed bya plugging pulse which arrives at the exhaust portjust prior to closure. These features maximise boththe airflow into the engine and the mass of thecharge trapped in the cylinder when the port closes.At 16000 rev/min, it is clear that the plugging pulsearrives after the exhaust port has closed.

Fig. 12 Exhaust Pressure at 10000 rev/min

Fig. 13 Exhaust Pressure at 16000 rev/min

STEADY STATE RESULTS - A comparison of thesteady state and acceleration results is presented inFig. 14. These results show that under steadyconditions, the bmep was up to 10% less than thatrecorded by the inertial dynamometer. It is believedthat a number of factors given below havecontributed to this result.

1. During steady state testing, the crankcasetemperature increased from 35°C to 50°C. Theplug seat temperature increased from 75°C to150°C. These temperature increases occurreddespite a cooling down period between eachmeasurement.

2. At the higher speeds, a relatively rich mixturesetting was used to prevent seizure of the engine.

3. The control of air/fuel ratio during the accelerationtest may differ from that of the steady statecondition.

4. The steady state exhaust gas temperatures maybe significantly different from those during anacceleration test.

In an effort to ascertain the influence of the finalfactor, a fast response thermocouple was installed inthe mid-section of the exhaust pipe. Thisthermocouple had a 0.5mm diameter inconel sheathand had a grounded junction to give improvedthermal conductivity. The actual responsecharacteristics of the thermocouple have not yetbeen measured, but in the near future, a single pulserig will be used to determine the response to a stepchange in gas temperature.

Fig. 14 Comparison of bmep Characteristics

The results presented in Fig. 15 show that, for theinertial dynamometer, lower exhaust gastemperatures were recorded between 7000 and12000 rev/min. For most of this speed range, thelower temperatures will reduce the pressure wavevelocity and will give improved exhaust tuningcharacteristics by delaying the arrival of the pluggingpulse. This feature may contribute to the measuredbmep difference. Between 13000 and 14000 rev/min,the temperatures are very similar, and thecorresponding bmep results also show little variation.Above 14000 rev/min, higher temperatures wererecorded on the acceleration test. These highertemperatures will increase the pressure wavevelocity and, in this speed range, improved theexhaust tuning effects will result.

The results recorded to date indicate that the twotest methods give significant differences in exhausttemperature which in turn produce marked changesin performance. Clearly, further investigation of theseexhaust temperature effects is necessary todetermine the nature and extent of thesephenomena.

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Fig. 15 Comparison of Exhaust Gas Temperatures

COMPUTER SIMULATION

Today, much of the understanding of physicalsystems is derived from computer simulations whichincorporate mathematical models for the relevantphysical processes. Development of these models isoften aided by experimental techniques which helpidentify the important variables, and either providevalidation for a theoretical analysis or a basis for anempirical relationship. Hence, a typical simulationmodel for any complex physical system maycomprise interacting theoretical and empirical sub-models which collectively predict the overallresponse to changes in design or external inputs.

In the case of most reciprocating engines, thebasic operating principles are easily understood, andsuccessful engines can be designed withoutrecourse to simulation techniques. However, the fluiddynamics and thermodynamics of the internalprocesses are extremely complex, and thus in manyinstances the use of ‘cut and try’ methods involvesexcessive experimentation and prototypedevelopment. The continual quest for techniqueswhich aid the rapid development of improvedengines has provided the stimulus for the study oftheoretical simulation techniques. At The Queen’sUniversity of Belfast (QUB), the computer modellingof internal combustion engines has been majorresearch topic for over twenty years. The majority ofthis research has been directed at two-strokeengines, but four-stroke, rotary and gas turbineengines have also been considered.

The two-stroke engine, although mechanicallysimpler, possesses a greater degree of fluid dynamiccomplexity than its four-stroke counterpart. Duringthe cylinder blowdown, scavenging and trappingphases, the flow into and out of the cylinder isprimarily due to the action of pressure waves, andnot cylinder volumetric changes, as in the case of thefour-stroke engine. Consequently, the performancecharacteristics are more markedly influenced bychanges in the dimensions of the inlet and exhaustducts, and the orientation of the ports.

For example, on a single-cylinder racing engine,replacement of a tuned exhaust pipe with a straightpipe of constant cross-section would result in adecrease in peak torque of at least 50%.

In general, the central element in the modelling ofreciprocating engines is the simulation of theunsteady gas flow in the inlet and exhaust tracts. Inthe case of the two-stroke engine, the accuratesimulation of unsteady gas flow phenomena isessential. In the exhaust duct, finite amplitudepressure waves result from the release of highpressure exhaust gas; in the inlet, pressure wavesresult from the volumetric changes in the crankcasevolume due to piston motion. Until recently, the‘Method of Characteristics’ (5) was primarily used asthe basis for simulation of the pressure wave action.However, an alternative method developed byBlair(6) is now employed. This non-isentropic methodhas been extensively validated on a pressure wavesimulator (7) and its claim to accuracy is increasinglysupported by correlation with measurements fromfiring engines.

In the simulation of the two-stroke engine, modelsare also required to simulate the cylinderscavenging, combustion, heat transfer and port flowprocesses. In addition, for reed valved two-strokeengines, a model is required to simulate the complexflow characteristics and dynamic action of the reedpetals under the action of gas pressure forcingfunctions. Extensive experimental studies (8,9,10)have led to the development of validated computermodels to simulate these elements of the engine’soperation.

KART ENGINE MODEL - A computer model ofthe 100cm3 kart engine was set up to elucidate thewave action phenomena and to aid the futuredevelopment of improved exhaust systems. Themodel included detailed geometrical and physicalinformation on all aspects of the engine ducting,cylinder, crankcase, combustion chamber, reedvalve, etc. For example, the model of the reed valveincluded 17 geometrical parameters in addition to thedensity and Young’s modulus of the reed petalmaterial. The exhaust system details used in thecomputer model were, those illustrated in Fig. 8 Asolid reverse cone was included as a validated modelfor the perforated variant is not yet completed.

The computer model predicts the pressure-timehistories, mass flow rates and temperaturesthroughout the engine at discrete crankshaft angleintervals. It is assumed that the engine is operating ina steady state condition and that a constant air/fuelratio is maintained. During the computation, agraphical display shows the piston position and thevariation of the major parameters during each enginecycle. After completion of each cycle, the modelgenerates data for the performance characteristics ofbmep, bsfc, delivery ratio, peak cylinder pressure,etc. Stability is attained after 10 to 20 cycles, and thesimulation then progresses to the next engine speed.

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Fig. 16 Measured and Calculated Delivery RatioCharacteristics

Fig. 17 Measured and Calculated bmepCharacteristics

SIMULATION RESULTS - The measured andcalculated delivery ratio characteristics are shown inFig. 16. Despite the wide range of speedsconsidered, reasonable correlation is evident. Thecorresponding bmep characteristics given in Fig. 17show excellent agreement. A sample plot showingmeasured and calculated exhaust pressure ispresented in Fig. 18. Comparison of the tracesshows that all the major features of the measuredtrace are reproduced by the calculations. Thepressures given in Fig. 18 represent the actualdynamic pressure as recorded by a transducer100mm from the exhaust port. This pressure is thesuperposition of leftwards and rightwards movingwaves which cannot be separated by anyinstrumentation. However, within the computermodel, these waves are calculated independentlyand graphical output showing each wave can beproduced. This feature of the model is often verybeneficial in elucidating the exhaust pressure waveaction.

Fig. 18 Comparison of Measured and CalculatedExhaust Pressures at 8000 rev/min

RECOMMENDATIONS AND LIMITATIONS

The inertial dynamometer system described inthis paper has proved to be a reliable and usefuldevice. However, in light of the experience gained indesigning and operating the dynamometer, a numberof possible improvements have become evident.Firstly, the shear pin has proved to be unnecessary,as seizure is extremely unlikely during the briefacceleration period. In the event of a seizure, theclutch can be used to disengage the flywheel.Furthermore, tensile tests on specimens from a kartchain has shown that the chain will break at arelatively low torque level.

The initial design for the dynamometer did notinclude the automotive clutch, which was added afterpreliminary tests showed difficultise in starting theengine. Consequently, the addition of the clutchassembly involved a number of design compromises.Clearly, the clutch friction plate could act directly onthe main flywheel, thereby eliminating the need forthe small secondary flywheel seen in Figs. 3 and 4.The design and manufacture of a more compactdynamometer is currently in progress.

The current limitations of the acceleration testmethod are the absence of any means to measureair-flow and air/fuel ratio.

CONCLUSIONS

The design and development of the inertialdynamometer system has proved to be an interestingand challenging project theme for mechanicalengineering students. The breadth of the project andthe high performance aspects have inspired thestudents, and excellent commitment and quality ofwork have resulted. When employed in the testing ofa high performance two-stroke engine, the inertialdynamometer has eliminated many of the difficultiesencountered during steady state testing, and hasprovided a test method which more closelyrepresents the conditions of the race track. The

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addition of a system for recording a series ofpressure traces at predetermined engine speedsduring a single acceleration has provided a powerfultool to aid the tuning of engines and the validation ofcomputer simulations.

ACKNOWLEDGEMENTS

The authors acknowledge The Queen’sUniversity for provision of the manufacturing,experimental and computing facilities employed inthis study. The contributions of QUB technicianRaymond McCullough, students Robin Montgomery,Clive Stewart, Paul McEntee, David Scullion, PeterTurner and Adrian Cunningham, and visitingstudents Kai, Stephano and Stephane are alsogratefully acknowledged. Castrol International Ltd. isacknowledged for supply of the test engine andlubricants.

REFERENCES

1. Karl-Heinz Bendix, ’Optimizing of thenonsteady operational behaviour of highperformance two-stroke engines’, SAE PaperNo. 931509, SETC Conference, Pisa, 1993.

2. Kart-Heinz Bendix, ’Optimierung desInstationaren Betriebsverhaltens vonHochleistungszweitaktmotoren’, 5thInternational Two-Wheeler Symposium,Technische Universitat Graz, Austria, April1993.

3. H. Holzleitner, ’Snowmobile 340ccm Formula 1- and Kart 100ccm Racing Engines: TwoExtremes in Terms of High-Performance Two-Stroke Engine Layouts’, SAE Paper No.911312, 1991.

4. G.P. Blair, The Basic Design of Two-StrokeEngines, SAE Publication No. R-104, ISBN 1-56091-008-9, 1990.

5. R.S. Benson, The Thermodynamics and GasDynamics of Internal Combustion Engines,Volumes 1 and 2, Clarendon Press, Oxford,1982.

6. G.P. Blair, ’An Alternative Method for thePrediction of Unsteady Gas Flow Through theInternal Combustion Engine’, SAE InternationalOff-Highway & Powerplant Congress,Milwaukee, Wisconsin, September 1991, SAEPaper No. 911850.

7. S.J. Kirkpatrick, G.P. Blair, R. Fleck and R.K.McMullan, ’Experimental Evaluation of 1-DComputer Codes for the Simulation of UnsteadyGas Flow Through Engines - A First Phase’,SAE International Off-Highway & PowerplantCongress, Milwaukee, Wisconsin, September1994, SAE Paper No. 941685.

8. M.E.G. Sweeney, R.G. Kenny, G.B. Swannand G.P. Blair, ’Single Cycle Gas TestingMethod for Two-Stroke Engine Scavenging’,SAE International Congress, Detroit, Michigan,February 1985, SAE Paper No. 850178.

9. M.G. Reid and R. Douglas, ’Quasi-Dimensional Modelling of Combustion in a Two-Stroke Cycle Spark Ignition Engine’, SAEInternational Off-Highway & PowerplantCongress, Milwaukee, Wisconsin, September1994, SAE Paper No. 941680.

10. R. Fleck, G.P. Blair, R.A.R. Houston, ’AnImproved Model for Predicting Reed ValveBehaviour in Two-Stroke Engines’, SAEInternational Off-Highway & PowerplantCongress, Milwaukee, Wisconsin, September1987, SAE Paper No. 871654.