200-01-3546 Racing Engine Design Options Investigated by Engine Simulation

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SAE TECHNICALPAPER SERIES 2000-01-3546

Racing Engine Design OptionsInvestigated by Engine Simulation

Gordon P. BlairMechanical Engineering Department

The Queen’s University of Belfast

Dermot O. MackeyOPTIMUM Power Technology

Reprinted From: Proceedings of the 2000 SAE MotorsportsEngineering Conference & Exposition

(P-361)

Motorsports Engineering Conference & ExpositionDearborn, Michigan

November 13-16, 2000

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2000-01-3546

Racing Engine Design Options Investigated byEngine Simulation

Gordon P. BlairMechanical Engineering Department

The Queen’s University of Belfast

Dermot O. MackeyOPTIMUM Power Technology

Copyright © 2000 Society of Automotive Engineers, Inc.

ABSTRACT

The paper discusses the design of a racing motorcycleengine to compete in World Superbike racing. This classof motorcycle racing is based on production machineswith four-stroke engines only. The rules allow threeengine variants to be used, a 750 cm3 four-cylinderengine, a 1000 cm3 twin-cylinder engine, and a 900 cm3

three-cylinder engine. To date only the first two variationshave been employed but this paper shows that the 900cm3 engine has the highest potential power output of theset. This is demonstrated using engine simulationsoftware and the finest detail of the design of the engineand its ducting are supplied within the discussion.

The input data for the engine simulation is provided byempiricism so that the design is initially well-matchedfrom the intake bellmouth to the end of the exhaustsystem. The outcome of this empirical process isconfirmed by the engine simulation to be a relevant initialdesign procedure.

1.0 INTRODUCTION

The racing in World Superbikes is among the hardest andmost closely fought in all of automotive sport andprovides a spectacle which has enthralled racingenthusiasts world-wide. It is not too much of anexaggeration to say that the Italian company of Ducatihas been the most successful over the years,campaigning a vee-twin engine of firstly 888 cm3, thenwith 916, 955 and finally with engines of 995 cm3

capacity. The design of the 955 cm3 variation has alreadybeen presented and debated at an earlier MotorsportsEngineering Conference by Boretti [1]* and is discussedin more detail by Blair [2]. In the early days of this racingclass, the 1000 cm3 vee-twin enjoyed a weight advantage

according to the rules established by the FédérationInternationale Motocycliste (FIM), but such was thecontinuing success of the Ducati that by 1998 allmachines had a minimum weight set at 162 kgirrespective of the engine type being employed. Some ofthe opposition finally paid Ducati the supremecompliment by opting for vee-twin engines for the year2000 racing series.

During all of this racing, none accepted the challenge tocompete with the third engine variant, namely the 900cm3 three-cylinder version, possibly because of the FIMhomologation requirements to have in production amotorcycle fitted with this engine type. Perhaps onlyTriumph in England, or Benelli in Italy, come close tosatisfying the rules in this regard. Also, it may be that theseveral manufacturers competing in World Superbikeracing considered that this engine configuration would benon-competitive. If that theory ever existed, thediscussion in this paper should dispel that opinion.

There is no doubt that a vee-twin engine provides themotorcycle with the narrowest frontal area profile, hencereducing the aerodynamic drag at high speed. Almost bydefinition, an across-the-frame four-cylinder unit has thelargest frontal area. Nevertheless, with a (750) vee-fourdesign, as in the RC45 design raced by Honda, there isvery little difference in the frontal area of the vee-twin andvee-four engine layouts. To put some flesh on thisdiscussion, the sketch in Fig. 1 is provided. It shows,drawn to scale using the cylinder bore dimensionsdeduced later and shown in Fig. 2, the basic frontalprofile exhibited by most of the logical cylinder layouts forthis racing class. While the primary gearbox and thecamshaft drives are not included in this sketch, Fig. 1, thebasic point regarding aerodynamic drag is well-illustrated. The 750 cm3 in-line four-cylinder unit is theworst-case drag scenario, the 1000 cm3 vee-twin is the

* Numbers in parentheses indicate References.




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Figure 1. Drag profiles of Superbike engine blocks.

best with the 750 vee-four being very similar, but the 900cm3 three-cylinder unit is not hugely worse-off in frontalarea than the vee-four and certainly better-off than the in-line four-cylinder case.

There is no detailed discussion here regarding enginemass, but, if one takes the block area profiles for eachengine as shown in Fig. 1 as being roughly indicative oftheir potential engine mass and bulk, it is not improbablethat the 900 cm3 three-cylinder unit could have the lowestvalues of any shown there.

If there is any future in World Superbike racing for a 900cm3 three-cylinder engine it must have at least equal

performance with the other machines currently racing. Inrecent times, Honda have raced 750 cm3 vee-four and,now, 1000 cm3 vee-twin machines. Kawasaki and Suzukirace with 750 cm3 in-line four-cylinder engines. Apriliaand Ducati use 1000 cm3 vee-twin power units. There aretwo ways to conduct this investigation into the potential ofthe 900 cm3 engine, either experimentally or purelytheoretically. Until recent times, the former, i.e., by designbased on experience, then building, cutting, trying andmodifying often proved to be a long-drawn out, expensiveand possibly fruitless process but was the only possiblesolution. The alternative is to investigate the potential ofthis, or any other, design theoretically using an accurate1D engine simulation code [5]. The word ‘accurate’ is anessential adjective to use here as an engine simulationmodel that is not an absolute predictor could provideequally confusing output information as the poorly-conducted experimental approach. The requirements forsimulation model accuracy are clearly spelled out by Blair[2], such as the employment of non-isentropicthermodynamics and gas dynamics and particle trackingthroughout the engine and ducting, branched pipemodels which include relevance to inter-pipe junctionangles, multi-zone combustion processes, and the use ofmaps of discharge coefficients based on both pressureratio and geometry for every type of boundaryencountered within the powerplant envelope. In theabsence of some or all of the above requirements, someengine simulation engine models need ‘calibration withexperimental data’ to be ‘accurate’. In the design contextbeing investigated here only the engine concept exists sosuch ‘calibration’ is not possible. Hence, an enginesimulation which is ‘absolute’ in its accuracy is essential.

In the theoretical design process which is to beconducted here, a second requirement is essential. Theneed is for the rapid assembly of all relevant data for theengine simulation, i.e., the data for the engine andducting geometry, including that for the valves, whichdata in its entirety should be selected so as to besufficiently well-matched in order to provide the requisiteengine breathing, tuning, and performancecharacteristics within the selected engine speed range.The need for this process, and the empiricism to supplyit, can be found in Reference [2].

2.0 THE 900 cm3 THREE-CYLINDER ENGINE

Target power and cylinder dimensionsThe assessment of the potential power output of anengine, including the basic cylinder dimensions toaccomplish it, is described elsewhere [2, Ch.1, Sec.1.8].Indeed, one of the engines used in illustration of theassessment technique is, no less than, a 1000 cm3

vee-twin racing motorcycle engine! The theory isprogrammed within a software package [3], and the input




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Figure 2. Engine geometry and performance data.

data to it for the 750 cm3 four-cylinder, 1000 cm3 twin-cylinder and 900 cm3 three-cylinder engines is illustratedin Fig. 2. The basic data needed is; total engine capacity;number of cylinders, bore-stroke ratio; mean pistonspeed; and the bmep expected at the engine speedwhere peak power is required. These numbers are basedon experience and, apart from the number of cylindersand the cubic capacity, they are used commonly for thethree candidate engines so as to give relevantcomparisons based on some expectation of equality ofbehaviour. The output data from the theory and thesoftware is also shown in Fig. 2. To the afficianado theoutput data for the first two Superbike engines for thebore, the stroke, and the peak power at the rated enginespeed, will seem familiar.

For the candidate 900 cm3 engine, the output dataestablishes the bore, the stroke and an engine speed of13500 rpm, at which speed it must run and breathe andburn so as to provide the 12.8 bar bmep (torque) if it is toattain the 173.5 bhp (129.4 kW) predicted for it. All otherthings being equal, if it can do that it will be the winner ofthe horsepower race by some 4.4 bhp. In short, it wouldnot only be competitive, it would be very competitive.The candidate engine has a bore of 82.5 mm and astroke of 56 mm. Simplistically, for the simulation analysisto follow, let a connecting rod of 105 mm centres beattached to it and a geometric compression ratio of 11 beapplied to each cylinder.

The valves in the cylinder headThere are many requirements for the design of the valvesin the cylinder head but priority must be given to makingthem fit within it! A conventional four-valve pent-roofcylinder head is selected as the design base and asoftware package [3] is used to select the size of thevalves. The output data, which shows the outcome of thisprocess for this 82.5 mm cylinder bore, is drawn to scalein Fig. 3. The input data specifies; the cylinder bore; theintake/exhaust valve size ratio; the radial side clearancesof the exhaust and intake valve from the cylinder bore,the inter-valve clearance between the intake valves; the

Figure 3. Cylinder head geometry and valve locations.

inter-valve clearance between the exhaust valves; andthe inter-valve clearance between the intake and exhaustvalves. The valve size in question is the outer seatdiameter of each valve. The input data, in mm, in theabove sequence is; 82.5; 1.14; 1.5 and 2.5; 2.4; 5.2; 3.0.The location of the valves in plan view is dimensioned inFig. 3. The outer seat of an exhaust valve is found to be28.5 mm and that of an intake valve is 32.5 mm.

To accommodate the necessary clearance volume of 30cm3 above a (more or less, apart from the valve cut-outs)flat-topped piston and with a squish clearance of 1.2 mm,then pent-roof head angles of 14 and 13 degrees, i.e., anincluded valve angle of 27 degrees, will be necessary;this is illustrated to scale in Fig. 3. The spark-plug, of 10mm, is centrally-located between the valves.

The basic dimensioning of each valveThe outer seat diameter, dos, of the exhaust valve is 28.5mm. As described by Blair [2, Ch.1, Fig. 1.14], a planseat width of 1.25 mm is selected, so the inner seatdiameter of the exhaust valve, dis, is 26.0 mm. Let themaximum exhaust valve lift, Lv, be 10.5 mm, i.e., 40% ofthe inner seat diameter.




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Figure 4. Cylinder head valve and manifold data.

The outer seat diameter, dos, of the intake valve is 32.5mm. Similarly [2, Ch.1, Fig. 1.14], a plan seat width of1.25 mm is selected, so the inner seat diameter of theintake valve, dis, is 30.0 mm. Let the maximum intakevalve lift, Lv, be 12.0 mm, i.e., also 40% of its inner seatdiameter.

Let the intake and exhaust valve stems, ds, each have adiameter of 6.5 mm. The ports at each valve have thesame dimensions as the inner seats, i.e., dp equal dis. Allof the above data, and that to be deduced below, aresummarised in Fig. 4.

The valve timingsIn Blair [2, Ch.6, Sec.6.1], there is empiricism providedfor the timing of valves of known physical dimensions soas to potentially give the requisite breathingcharacteristics at rated speed and load. Valve liftcharacteristics are also discussed in the same Reference[2, Ch.1, Sec.1.5] and the ensuing lift curve determinedin terms of ramp lift ratios with the upshot calculated interms of lift, velocity and acceleration with respect tocrankshaft angle at any given engine speed. Thesetheories are programmed into software [3] and usedhere. The engine speed for the computations is at 13500rpm, the engine speed where peak power is anticipated.

For the intake and the exhaust valves, the ramp lift ratio,Cr, is 0.20, i.e., it is 20% of the maximum valve lift andoccurs over a 40 [°crank] ramp duration.

The outcome of using the software is that the followingvalve timing events are selected; the valve opening andclosing events, vo and vc, are 60 °btdc and 87 °abdc forthe intake valve, while they are 90 °bbdc and 60 °atdc forthe exhaust valve. These selections are based on thegeometry described above and the specific time-areasfound by the programmed theory for these valves to be inaccordance with the empiricism specified in theReference [2, Ch.6, Sec.6.1]. The time-area analysis forthe intake valves shows that the pumping and rammingperiod values have specific time-area values of 129.6 x10-4 and 17.56 x 10-4 s/m, respectively. The time-areaanalysis for the exhaust valves shows that the exhaustblowdown and pumping periods have specific time-areavalues of 14.7 x 10-4 and 97.3 x 10-4 s/m, respectively.

The exhaust and intake overlap valve periods havespecific time-area values of 29.5 x 10-4 and 39.7 x 10-4

s/m, respectively. All of these empirical criteria suggestthat these valves, opened, closed and lifted at thesetimings and rates at 13500 rpm would breathe sufficientlywell to provide 12.8 bar bmep, which also assumes thatall other tuning and duct flow criteria are to be designedto match these valves.

For the record, at 13500 rpm, the maximum valve velocityfor the intake valves is 11 m/s and the maximumacceleration is 3400 g. For the exhaust valves, theequivalent maxima are 9 m/s and 3000 g. This valveand lift geometry should be capable of being handledby a normal double-overhead camshaft system withconventional coil springs for each valve.

The manifolds at the cylinder headAs mentioned above, unless the ducting system is tunedto match the valve timing and geometry then the outcomewill be poor cylinder filling and emptying. The first gambitin this argument is the size of the aperture at the cylinderhead to which the exhaust and intake ducting isconnected for this proportions the maximum amplitude ofthe intake and exhaust waves proceeding into their ducts.This topic is discussed at great length in the Reference[2, Ch.6, Sec.6.1.9] in terms of the requisite manifold-portarea ratios, Cm, for the exhaust and intake system at thecylinder head.

For the exhaust system the value of the manifold-portarea ratios, Cm, is normally in the range 1.2 to 1.4, sothat the maximum amplitude of the exhaust pulse whichis created at the maximum power engine speed is about2.0 bar. Any larger exhaust pulse will tend to steep-frontthe exhaust blowdown pressure wave and lose waveenergy in the process and a lesser amplitude exhaustwave will provide weak tuning from its wave reflectionswithin the exhaust system. From Fig. 4, it can be seenthat a value of 1.26 is selected, which devolves to anexhaust manifold diameter, dm, of 40 mm at the cylinderhead face.

For the intake system the value of the manifold-port arearatios, Cm, is normally in the range 0.9 to 1.0, so that theamplitude of the intake pulse which is created is as deepa suction wave as possible without incurring energylosses in its wave transmission to the bellmouth. Thedeepest suction wave provides the largest ramming wavereflection at the bellmouth end which is absolutelynecessary so as to provide airflow Delivery Ratios inexcess of unity. From Fig. 4, it can be seen that a value of0.93 is selected, which devolves to an intake manifolddiameter, dm, of 40 mm at the cylinder head face.

The coincidence of the two numbers, i.e., that the intakeand exhaust manifold diameters are both 40 mm, isunexceptional but should not be taken as a ‘rule ofthumb’!




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Figure 5. Exhaust and intake ducting layout and data.

The engine ductingA dimensioned sketch of the engine ducting is shown inFig. 5. It is not to scale.

The basic layout of the exhaust system is a ‘three-into-one collector’ system which, unlike car racing engines inFormula 1 or IRL, cannot end in the atmosphere. Allmotorcycle racing is conducted with (partially) silencedengines which must meet FIM-dictated noise legislationand testing conducted by FIM Technical Stewards. Unlikemuch automobile racing, attending motorcycle racing isan aural pleasure and not pain! Hence, at the end of thecollector pipe a small primary plenum is installed toprovide a ‘dummy atmosphere’ to induce the necessarywave tuning characteristics of a 3-1 collector pipesystem. Beyond that, two further pipes with a combinedarea equal to that of the 80 mm collector pipe proceed totwo mufflers and tailpipes and thence to the atmosphere.The fundamental design thinking is to place the primaryexhaust plenum below and behind the engine/gearboxunit and in front of the rear wheel of the motorcycle andthen route either side the final pipes to two silencersplaced under the rider’s seat on the motorcycle, à laDucati.

The basic layout of the intake system is to connect thethree intake ram pipes, which have bellmouth ends, to anairbox plenum situated behind the steering head andbelow the fuel tank, from which airbox a duct goesforward to the front of the motorcycle fairing to collect airpressurised by the forward motion of the machine. In theengine simulation analysis conducted here that forwardmotion is presently neglected so as to ascertain thereference engine performance characteristics in still air,i.e., as on the dynamometer during performance testing.

The exhaust system designThe basic design of the primary pipe and collector-pipeexhaust systems is described elsewhere [2, Ch.6, Secs.6.4.5-6]. The empirical theory detailed there is

programmed into software [3] and used here with acollector area ratio, Ccoll, of 6.0. To optimise the systemtuning at 13500 rpm, the prediction is that the primarypipe length should be 450 mm, and the collector tailpipeshould be 640 mm long with a diameter of 80 mm. Theprimary pipe length of 450 mm is split between 75 mmwithin the cylinder head and 375 mm beyond it to thecollector. At the four-way branch the inter-pipe angles areeach assumed to be a mutual 10 degrees to the outletpipe axis.

The remainder of the piping system with the silencersthen mimics this 640 mm tuning length with two pipeseach of 56 mm diameter which equates outflow from theprimary plenum to inflow from the 80 mm diameter of thecollector pipe. The above dimensions may be foundsketched in Fig. 5.

The intake system designThe main design item here is the length for the intakevalves to the bellmouth end so as to optimally phase theramming at the engine speed for maximum power,namely 13500 rpm. This issue is extensively discussedelsewhere [2, Ch.6, Sec. 6.3]. The empirical theorydetailed there is programmed into software [3]. Theoptimum ramming length for the intake tract is found to be230 mm and this dimension can be observed sketched inFig. 5. The empirical theory also warns us that while theintake ramming is set to peak at 13500 rpm, and thatthere will be another ramming peak at 10000 rpm, therewill also be troughs in the engine airflow curve caused byineffective ramming at 11500 rpm and at 8700 rpm. Thislatter speed of 8700 rpm is of little concern for a racingengine where an effective engine speed range of 3000rpm is more than adequate, but the warning of a troughat 11500 rpm is of much greater concern so the effect ofthat will be closely monitored within the results obtainedfrom the engine simulation.

The minimum effective airbox size is some 5 to 6 timesthe total cylinder capacity, so an airbox of some 5 litres isseen sketched in Fig. 5. In practice, it could well be(usefully) larger. In this first instance, the intake duct tothe atmosphere is set large and short, but it should bethe subject of further optimisation as an important tuningissue has been shown to be involved [2, Ch.7, Sec.7.6.5].

Data is assembledHaving assembled all of the geometric data for thesimulation of the engine, this data is presented to theengine simulation [5] for the computation and subsequentanalysis of the calculated performance characteristics.

3.0 THE 900 cm3 ENGINE SIMULATION

The simulation modelThe simulation model employed is the professionaledition of VIRTUAL 4-STROKE [5] by comparison withthe SAE edition associated with Blair [2]. The SAE edition




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Figure 6. Power output.

of VIRTUAL 4-STROKE [4] handles only a single-cylinderengine but is otherwise similar in operation to theprofessional edition which can handle any number ofcylinders inter-connected, or not, by intake or exhaustducting.

Combustion within the simulation modelThis is as described in Reference [2] and, in the absenceof precise burn data for this mythical 900 cm3 engine andcombustion chamber, experimentally-derived data foranother 4-valve pent roof chamber is ascribed here withinthe simulation model. The precise data can be foundwithin this same Reference [2, Ch.4, Fig. 4.10].

It is assumed within the simulation that the engine willburn unleaded fuel with an air-fuel ratio of 12.0.

Discharge coefficients within the simulation This very important issue is well-debated by Blair [2] andthe discharge coefficient, Cd, maps for the valves in thisengine are also experimentally-derived from a 4-valvepent-roof racing engine cylinder head and shownelsewhere [2, Figs. 3.37-3.40] for both directions at theintake and exhaust valves. The discharge coefficientmaps for the bellmouth or plain ends of pipes at theatmosphere or a plenum can be found from the samesource [2, Figs. 3.6-3.10].

Further relevant simulation dataIt is assumed that the engine breathes to and from a(static) atmosphere at 1.01325 bar and 25 °C.

It is also assumed that the in-cylinder surfacetemperatures are 300, 200 and 150 °C for the cylinderhead surface, the piston crown surface, and the cylinderliner, respectively. A (incomplete) list of the pipe skin

Figure 7. Torque output.

temperatures are as follows; intake duct, 200 °C withinthe cylinder head and 35 °C beyond it; primary exhaustduct, 250 °C within the cylinder head and 450 °C beyondit; collector tail pipe and exhaust tail pipes, 400 °C.

4.0 THE 900 cm3 ENGINE SIMULATED

The simulated performance characteristicsThe overall performance characteristics predicted by theVIRTUAL 4-STROKE [5] simulation are shown in Figs.6-13. They are plotted in both Imperial and SI units towiden the relevance of the discussion.

Power outputPower output is of primary importance in racing enginedesign. It can be seen in Fig. 6 that the engine has apeak power of 172 bhp (128 kW) phased as required at13500 rpm. This corresponds closely with the requiredpower for the target design seen in Fig. 2. The powerspread is seen to be satisfactorily wide over a 3000 rpmpower band with 149 bhp (111 kW) available at 10500rpm. While there is no dip in the power curve at 11500rpm, as warned by the empirical intake ramminganalysis, it is clear that the power curve does increasemore steeply after that engine speed.

Torque and brake mean effective pressure (bmep)This is shown in Figs. 7 and 8. In Fig. 8 the peak torque isat 10500 rpm (or below) at 101 Nm (75 ft-lbf) and it canbe observed that there is a torque dip at 11500 rpmbefore rising to a second peak at 12000 rpm.

The bmep profile in Fig. 8 naturally mimics the torquecurve, with a peak at 10500 rpm of 14.2 bar beforeyielding a second peak at 12500 rpm of 13.3 bar. At the




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Figure 8. Brake mean effective pressure.

peak power engine speed of 13500 rpm the bmep is12.70 bar (187.6 lbf/in2) which is almost identical to thetarget value of 12.8 bar seen and set in Fig. 2.

Brake specific fuel consumption (bsfc)The bsfc profile for the engine is shown in Fig. 9 and itcan be seen that values in the peak power zone areabout 305 g/kWh (0.5 lb/hp.hr). Although this matter hasnot been aired before it is a fairly important issue as theraces in World Superbike racing are of a notinconsiderable length. Most legs, there are two legsraced each day at any given Championship meeting, lastfor about forty minutes. Taking the average power usageat 60% of maximum, and using the above bsfc figuresover a 40 minute period, simple arithmetic yields a totalfuel consumption of 14 kg (31 lb). Assuming a specificgravity for the fuel of 0.8, that implies a possible total fuelconsumption over a 40 minute race of 17.5 litres (3.85Imp. Gall., 4.6 US gall.). If the brake specific fuelconsumption could be lowered by, for instance, 10%, thenthat reduces the potential top hamper of the motorcycleby 1.4 kg (3 lb) in mass terms and the bulk by 1.75 litres(0.38 Imp. gall.). Of these two reductions, the bulk is themore important as the airbox and fuel tank vie for thesame available space within the motorcycle frame and1.75 litres of space added to the intake airbox plenumwould increase it by 35%, undoubtedly improving theengine breathing and power performance.

Airflow as Delivery RatioThis is shown in Fig. 10, together with Charging andTrapping Efficiency [2]. It can be seen that peak airflow isindeed at the required speed for peak power at 13500rpm. Here the Delivery Ratio (DR) is commendably highat 1.14. The ramming length, allied to the valve geometryselected, is indeed phased correctly; the empiricismemployed [3] has been most effective in this regard. Itspessimism regarding the trough in the ramming process

Figure 9. Brake specific fuel consumption.

at 11500 rpm is seen to be justified, as too is its moreoptimistic prediction that another ramming peak exists at10000 rpm. Later, in this paper, the pressure (and other)time histories will be examined to confirm the origins ofthe positive tuning events at 13500 rpm and the negativetuning effects at 11500 rpm.

The Charging Efficiency (CE) and Trapping Efficiency(TE) show the overall outcome of much of the intake andexhaust tuning. It may be recalled that CE is the multipleof DR and TE and represents the dimensionless mass ofair trapped in the cylinder during each cycle. Hence, asthe CE profile is relatively flat with a maximum of 1.095and a minimum of 1.05, so too must be the profile of theindicated torque. It is clear that the trapping efficiencybehaviour offsets some of the effects of the poorramming at 11500 rpm; here the TE is 100%. However, it

Figure 10. Delivery Ratio and charging efficiencies.




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Figure 11. Indicated and pumping mep.

deteriorates the CE value at the peak engine speed asthe DR value is 1.14 but the TE value is 0.94. The originsof this behaviour can only be observed from the time-related events around the cylinder at the significantengine speeds and these are examined later.

Indicated and pumping mean effective pressureThe indicated mean effective pressure (imep), which isdirectly related to indicated torque, is shown in Fig. 11.The correspondence of the imep profile with that ofCharging Efficiency in Fig. 10 and which is explainedabove as being logical, is now clear.

The pumping loss (pmep) is recorded as pumping meaneffective pressure in the cylinder over the nominal opencycle, i.e., from +180 to +540 degrees crankangle, where0 degrees crankangle is the tdc point during combustion.The profile of pmep with respect to engine speed is quitenormal for a racing engine where it increases almostlinearly, and quite considerably, with that speed. Atmaximum engine speed it consumes some 5.5 % of thework done on the piston, i.e., the imep.

Friction mep and mechanical efficiencyThe characteristics for the engine friction loss, as frictionmean effective pressure (fmep), are an empiricalrelationship and the one used can be found elsewhere [2,Ch.5, Sec.5.3]. It is plotted in Fig. 12 and its assumedlinear characteristic is clear. For this engine it varies from3.25 bar to 3.95 bar over the speed range. As the brakemean effective pressure (bmep) is the residue of theimep after the pumping and friction loss (as pmep andfmep) are subtracted, the origins of the decaying bmepprofile in Fig. 8, from the relatively flat imep profile inFig. 11, are now obvious.

Figure 12. Friction mep and Mechanical Efficiency.

The increasing severity of the parasitic losses of pumpingand friction with respect to speed can also be seen in Fig.12. Here the Mechanical Efficiency (ME) of the engine isplotted and it has an almost linear profile of decay withengine speed. It will be recalled that ME is defined as thedivision of bmep by imep. At peak power it is 0.73 whichmeans that 27% of the energy provided to the piston hasbeen dissipated; but then all such Philistines will attestthat if the engine is merely used for racing it is supremelyobvious that 100% has been dissipated!

Noise outputMotorcycle racing engines have to meet FIM-dictatednoise legislation. The VIRTUAL 4-STROKE simulationcode [2, 5] predicts the noise level and noise spectra at

Figure 13. Exhaust and intake noise levels.




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Figure 14. Airflow at 11500 and 13500 rpm.

any point in space beyond the ends of the ducting. Theprofile of the intake and exhaust noise is theoreticallydeduced at a microphone location of 1.0 m in free spaceequidistant from the atmospheric exit of the intake andexhaust systems. The values of those noise levels overthe engine speed range are shown in Fig. 13 and areplotted in dB (A scale) units, i.e., as dBa. The exhaust isat its most raucous at the lowest engine speed, whereasthe intake system noise increases with engine speed.Curiously, the engine speed with the lowest intake airflow,i.e., 11500 rpm, is one of the loud zones in the intakenoise profile.

The total noise output, i.e., the combined intake andexhaust noise, is also plotted in Fig. 13. It is, of course, alogarithmic addition procedure.

The simulation does not debate the quality of the noisealthough it is possible to have the simulation playback theexhaust or the intake noise or the total noise spectra. Todate in World Superbike racing, the vee-twin engines withtheir off-beat note have been the machines to excite theears of the enthusiasts. This three-cylinder racing engineshould easily assume that role, as can be attested by anywho ever had the good fortune, or are old enough, toremember Giacomo Agostini and his MV Agusta!

5.0 ENGINE TUNING AT 11500 AND 13500 RPM

GeneralHinted at by the empiricism, it is shown above bysimulation that the engine breathes easily and is welltuned at 13500 rpm, but some deterioration of that tuningtakes place at 11500 rpm. The Delivery Ratio is 1.05 at11500 rpm but is 1.14 at 13500 rpm; see Fig. 10. Thesimulation [5] output can be queried at any pointthroughout the entire engine and its ducting to providetime-related data for pressure, temperature, mass flowrate, charge purity, density, etc., etc.

Figure 15. Pressures at 13500 rpm.

To illustrate this point, the dynamic growth of air flow (asDR) within cylinder number 1 is plotted in Fig. 14 withrespect to crankangle at 11500 and at 13500 rpm. Theyare only plotted over a relevant crankangle period tomake more obvious to the reader the DR profiles duringinduction. The valve event timing marks of intake valveopening and closure (IO and IC) and exhaust valveclosure (EC) are marked on the diagram. It should benoted that with symmetrical cylinder timing and identicalpiping for each cylinder the behaviour in the other twocylinders is similar to that in cylinder no.1; this is notalways true and can have deleterious effects, e.g., withasymmetrical cylinder timing events [2, Ch.5, Secs.5.7-8]. In Fig. 14 at 13500 rpm, it can be seen that theairflow into the cylinder commences almost as soon asthe intake valve opens. The DR reaches 25% by the timethe exhaust valve closes at 60 °atdc and then rises to amaximum of 1.14 by 40 °abdc and does not deteriorateuntil the valve closes. On the other hand, at 11500 rpm,the incoming airflow commences at tdc, and isdisadvantaged by some 10% at the exhaust valve closingpoint. While the profile apparently recovers in the middleof the suction stroke, that 10% disadvantage has notbeen recovered by 40 °abdc with some small backflowevident until the intake valve closes. How do these effectscome about at these engine speeds?

Pressure diagrams at 13500 and 11500 rpmIn Figs. 15 and 16 are plotted the pressures in thecylinder, and in the intake and exhaust ducts near to theintake and exhaust valves, at 13500 and 11500 rpm,respectively. The pressure oscillations at 13500 rpm arenoticeably more pronounced than at 11500 rpm,although the exhaust pressures around tdc in the overlapvalve period are somewhat similar at both 11500 and13500 rpm; here the collector pipe system has beentuned correctly to provide suction waves of some 0.6-0.7atm to assist with scavenging the cylinder of exhaust gaswith the intake valves open.




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Figure 16. Pressures at 11500 rpm.

It is the behaviour of the residual waves in the intakemanifold which causes the airflow differences at 11500and 13500 rpm. At 13500 rpm, a residual pressureoscillation in the intake tract exceeds the cylinderpressure from intake valve opening until exhaust valveclosure. As the exhaust pressure is less than either ofthem so flow takes place from intake through the cylinderto the exhaust pipe. At 11500 rpm, the residual pressureoscillation in the intake tract arrives too early and thecylinder pressure basically equals or exceeeds intake linepressure until tdc and after that air flows into the cylinder.However, the exhaust pressure is higher than either ofthem after tdc so exhaust gas flow takes place to thecylinder from the exhaust pipe.

Near intake valve closure, there is a higher rammingpressure in the intake tract at 13500 rpm, whereas that at11500 rpm arrives a little too early for perfection. Both ofthese primary ramming waves are quite effective; it is theweakness of the secondary ramming effect during the

Figure 17. Gas particle velocities at 13500 rpm.

valve overlap period which is the root cause of thereduced airflow at 11500 rpm [2, Ch.6, Sec.6.3]

Velocity diagrams at 13500 and 11500 rpmIn Figs. 17 and 18 are plotted the gas particle velocities inthe intake and exhaust ducts near to the intake andexhaust valves, at 13500 and 11500 rpm, respectively.The gas particle velocity is plotted as Mach Number (NBthe local speed of sound is about 350 m/s or 1150 ft/sec).

The remarks passed above regarding particle flow canbe seen to be confirmed. In Fig. 17, at 13500 rpm, theintake flow commences at intake valve opening and onlya negligible amount of backflow takes place at intakevalve closure. In Fig. 18, at 11500 rpm, the intake flowshows reverse flow from the cylinder until tdc and thenfurther loss of cylinder charge by reverse flow at intakevalve closure. There is reverse flow of exhaust gas intothe cylinder from about tdc until exhaust valve closure.This exhaust gas then occupies cylinder space which cannever be filled by intake air.

Charge purity at 13500 and 11500 rpmIn Figs. 19 and 20 are plotted the charge purities in thecylinder, and in the intake and exhaust ducts near to theintake and exhaust valves, at 13500 and 11500 rpm,respectively. It will be recalled that air has a charge purityof 1.0 and exhaust gas a charge purity of 0.0. The plotexamines only the critical valve overlap period.

The upshot of the pressure and velocity flow effectsaround the cylinder at these two engine speeds can nowbe observed in the quality of the gas which resides inthese locations at various junctures during this period ofengine operation.

In Fig. 18, at 13500 rpm, as there is no blowback into theintake tract due to the well-phased intake tuning, thecharge purity remains at unity throughout. The cylinderpurity rises from zero almost as soon as the intake valveopens and is 95% fresh air by the time that the exhaustvalve closes; the exhaust tuning has extracted fresh

Figure 18. Gas particle velocities at 11500 rpm.




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Figure 19. Charge purities at 13500 rpm.

charge right through the relatively small cylinder volume;at +/- 60 degrees either side of tdc the cylinder volume isnot much larger than the clearance volume. Naturally, thisexhaust gas extraction and cylinder scavenging processcomes at a price, namely the reduction of trappingefficiency at 13500 rpm as seen in Fig. 10.

By contrast, in Fig. 20, at 11500 rpm, the cylinder purityhovers around zero until tdc, then rises to about 82% atexhaust valve closure. That amount of exhaust gas ispermanently trapped in the cylinder. The intake systemrecords the exhaust blowback and falls to about 0.8before recovering as normal intake flow resumes, the firstinflow of which is the return of that same cylinder gas.The exhaust system sees very little exhaust systemtuning effect on cylinder scavenging and its fresh chargecontent barely gets to 4% by exhaust valve closure. Thisexplains the high trapping efficiency at 11500 rpm seenin Fig. 10.

Figure 21. Temperatures at 13500 rpm.

Figure 20. Charge purities at 11500 rpm.

Temperature at 13500 and 11500 rpmIn Figs. 21 and 22 are plotted the charge temperature inthe cylinder, and in the intake and exhaust ducts near tothe intake and exhaust valves, at 13500 and 11500 rpm,respectively. The plot looks only at the critical valveoverlap period.

The upshot of the pressure, charge purity and velocityflow effects around the cylinder at these two enginespeeds can now be observed in the temperature of thegas which resides in these locations at various juncturesduring this period of engine operation. Simplistically,exhaust gas is ‘hot’ and air is ‘cold’.

In Fig. 21, at 13500 rpm, the local gas temperature in theintake tract, due to local heat transfer and compressioneffects is about 60 °C and then drops as cooler air comesalong from the atmospheric end of the intake tract. Thecylinder temperature drops from over 700 °C to about

Figure 22. Temperatures at 11500 rpm.




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Figure 23. Performance re collector pipe diameter.

20 °C by exhaust valve closure due to the rapid andcontinuing inrush of cold air into the small clearancevolume. The exhaust temperature drops similarly asmuch fresh air is extracted right through the chamber intothe exhaust pipe due to the suction tuning effect of theexhaust pipe, aligning itself with the previously notedcharge purity effects in Fig. 19 and the trapping efficiencyeffects recorded in Fig. 10.

In Fig. 22, at 11500 rpm, the reverse flow of cylinder gas,i.e., exhaust gas, causes the temperature in the intaketract near to the valve to rise to about 160 °C. Thecylinder temperature drop basically stalls until after tdc,as no fresh air enters until that point and what then doesenter is neither pure air nor is it cool. The cylindertemperature falls after tdc, but its profile is flatter as hotexhaust gas is reverse flowing into the cylinder untilexhaust valve closure; see Fig. 18. The exhaust gastemperature in the duct remains high locally as very littlefresh air gets extracted into it that location.

Conclusions re 11500 and 13500 rpmThe analysis of these dynamic effects reveals the originsof the positive tuning at 13500 rpm, and the negativetuning at 11500 rpm. By exhaust valve closure as the realsuction process gets under way, at 11500 rpm bycomparison with 13500 rpm, the cylinder is already filledwith some 15% more exhaust gas and is 80 °C hotter.This is due to poorly-phased secondary intake ramming.The primary ramming process is also somewhat weakerand also relatively poorly phased, so further cylinder aircharge is lost near intake valve closure at 11500 rpm.The exhaust system is optimally phased at 13500 rpmand is still quite well tuned at 11500 rpm.

6.0 THE COLLECTOR PIPE DESIGN

GeneralIn the book by Blair it is attested that the design of acollector pipe system is best conducted with a collectorarea ratio, Ccoll, of 6.0 [2, Ch.6, Sec.6.4.6]. As this piece

of empiricism was culled from an analysis of ‘4-into-1’collector pipe systems and all of which collector tailpipesended in the atmosphere, this may not apply so readily tothis particular design of a three-cylinder engine wherethe tail-pipe ends in a small ‘pseudo-atmosphere’ plenumand is followed by further exit exhaust pipes andsilencers. The empirical use of a value of Ccoll of 6.0 forthis 900 cm3 three-cylinder engine yielded a collectortailpipe diameter of 80.0 mm, as shown in Fig. 5. Todetermine if the empiricism is still valid for a differingnumber of cylinders and for a revised final exhaust pipelayout, the collector tailpipe diameter is changed over afairly wide range from 74 to 92 mm and the VIRTUAL4-STROKE simulation rerun at 13500 rpm in each case.The range of collector area ratios examined is from 5.4to 7.3.

Effect of tailpipe diameter on powerThe result of conducting this engine simulation process isshown in Fig. 23. The power output varies over a narrowrange of just 0.5 bhp on 172 bhp, implying that collectortailpipe diameter is not important. However, powermaxima are observed around Ccoll values of 6.0 and 6.8,thereby justifying the use of the 6.0 criterion used for thedesign of this system. While the Ccoll value of 6.8 yields amarginally higher maximum power value, installing an88 mm diameter pipe underneath the engine/gearboxunit would certainly pose more problems than would the80 mm pipe designed using the 6.0 criterion.

Unsteady flow effects at the exhaust collectorBlair [2] gives a detailed explanation of the dynamicpressure and mass flow behaviour occasioned by anexhaust collector pipe system. It will be useful andinstructive from a design standpoint to determine if thereare behavioural differences within this system bycomparison with those examined previously in Reference[2].

At 13500 rpm with the exhaust system design as shownin Fig. 5, the predictions of pressure and mass flow withrespect to crankangle are extracted from the simulation.These are plotted in Figs. 24-27 for this three-cylinderengine. They are precisely analogous to, and may bedirectly compared with, those provided in the Reference[2, Ch.6, Figs. 6.69-6.72] for a four-cylinder engine withan equivalent collector exhaust system. There are twophysical locations where simulation data is collected, oneis just upstream of the collector in the pipe emanatingfrom cylinder number 1 and the other is just downstreamof the collector in the tailpipe.

In Fig. 24 is shown the superposition pressures eitherside of the collector. The superposition pressure is the‘addition’ of the leftward and rightward wave pressuresand is the pressure that one records with a pressuretransducer. This diagram is almost indecipherable interms of detailing the dynamic flow behaviour at thecollector, as already pointed out in the Reference [2,Fig. 6.69].




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Figure 24. Superposition pressures at exhaust collector.

In Fig. 25 are plotted the leftward and rightward wavepressures in the exhaust pipe from cylinder number 1.This picture is almost identical with that shown previouslyin the Reference [2, Fig. 6.70], indicating that thefundamental wave reflection behaviour in this collectorpipe system is the same. The exhaust pressure wavecoming from the engine, the rightward wave, has anamplitude of 2.1 atm, thereby justifying the initial designselection of the exhaust manifold to port area ratio, Cm, at1.26. The leftward wave, so necessary for good cylinderscavenging, is a suction wave of some 0.55 atm and it ispleasing to note that it almost identical to that describedfor a four-cylinder engine where the collector tailpipeended in the atmosphere [2, Fig. 6.70]. In short, thetuning action of the collector pipe system is notdiminished by the addition of the exhaust plenum and thefurther exhaust pipes and silencers. Indeed, thepresence of the plenum is vital in providing the dynamicsimilarity and the tuning action of the collector system forthis three-cylinder engine.

Figure 26. Individual wave pressures in the tailpipe.

Figure 25. Individual wave pressures in primary pipe.

In Fig. 26 are plotted the leftward and rightward wavepressures in the collector tailpipe. This picture isanalogous, and dynamically similar, to that shownpreviously [2, Fig. 6.71]. In the three-cylinder case thereare six oscillations on each pressure trace, bycomparison with the four oscillations for the four-cylinderengine. It may be presumed that the extra threeoscillations in Fig. 26 are generated by pressure waveaction in the final tailpipes and silencers.

In Fig. 27 are plotted the mass flow rates in the exhaustpipe from cylinder number 1 and in the collector tailpipe.This graph is almost identical to that shown earlier for a4-cylinder engine [2, Fig. 6.72], apart from the fact thatthere are now six mass flow oscillations in the tailpipe. Asfar as the action in cylinder number 1 is concerned it isidentical, which indicates that the extraction processgenerated by the collector tailpipe and plenum ensuresthat exhaust gas is extracted from each primary exhaust

Figure 27. Mass flow rates in primary and tailpipe.




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pipe, and hence that cylinder, without sending awkward(positive) pressure wave reflections up each primary pipeto interfere with the scavenging process at each cylinder.

7.0 CONCLUSIONS

The 900 cm3 three-cylinder motorcycle engineThe output data from the simulation indicates that thisengine would produce in excess of 170 bhp at 13500 rpmwith a fairly broad power curve, in which case it would bevery competitive in World Superbike racing. The engine isquite compact in mass and bulk terms making it an evenmore attractive prospect in this regard. With an in-lineconfiguration, the engine mass is preferentially located tooptimise the weight distribution within the motorcycle bycomparison with vee-engine designs.

The empirical design procedures described by Blair [2]for cylinder dimensioning, valve geometry, and the intakeand exhaust systems are shown to be relevant and toprovide a well-matched initial design prior to furtheroptimisation with an accurate engine simulation.

An accurate engine simulation [2, 5] provides invaluabledesign insights into the operational behaviour of theengine at any speed or load, be it about intake ramming,cylinder scavenging, or exhaust collector systemdynamics, etc. Without such insights, the development ofan ‘unknown’ racing engine by purely experimentalmeans is a frustrating, time-consuming, expensive, andpotentially fruitless exercise.

Exhaust system designFrom the evidence presented, it appears that the designapproach suggested in Reference [2] for the collector-type exhaust system of a high performance engine hasfairly wide applicability and can be extended to thosesystems which must contain silencers.

8.0 REFERENCES

1. A.A. Boretti, G. Cantore, E. Mattarelli, F. Preziosi,“Experimental and Computational Analysis of a HighPerformance Motorcycle Engine”, SAE MotorsportsEngineering Conference, Dearborn, MI, November1996, SAE 962526.

2. G.P. Blair, “Design and Simulation of Four-StrokeEngines”, Society of Automotive Engineers, R186,ISBN 0-7680-0440-3, Warrendale, PA, August 1999.

3. G.P. Blair, “Education and Design Software for Four-Stroke Engines”, Society of Automotive Engineers,R186SW, Warrendale, PA, August 1999.

4. G.P. Blair and OPTIMUM Power Technology,“VIRTUAL 4-STROKE Engine Simulation Model forFour-Stroke Engines”, Society of AutomotiveEngineers, R186M, Warrendale, PA, August 1999.

5. OPTIMUM Power Technology, “VIRTUAL 4-STROKE,Professional Edition”, Internet, www.optimum-power.com.




Documents

200-01-3546 Racing Engine Design Options Investigated by Engine Simulation