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Experimental analysis of a motorbike high speed racing engine
Applied Energy 87 (2010) 1641–1650
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Experimental analysis of a motorbike high speed racing engine Massimo Masi *, Andrea Toffolo, Marco Antonello Department of Mechanical Engineering, University of Padova, Via Venezia 1, 35131 Padova, Italy
a r t i c l e
i n f o
Article history: Received 27 April 2009 Received in revised form 25 September 2009 Accepted 25 September 2009 Available online 25 October 2009 Keywords: High speed engine testing Sweep test reliability Discharge flow coefficients
a b s t r a c t Power gain is the main objective in any motorbike competition. Despite of the wide literature on theoretical and experimental methods for increasing engine power, there is a general lack of data about tests on racing engine performance due to the obvious manufacturers’ reluctance to spread information, especially for recent high technological level applications. This paper, instead, presents all the main results of the experimental tests conducted on a motorbike engine both in the original stock arrangement and in a modified configuration proposed in compliance with the Technical Regulations of the 2007 FIM Road Racing Supersport Italian Championship (CIV). Traditional testing techniques (steady-flow discharge coefficients measurements and chassis dynamometer tests performed in the slow speed ramp mode) are chosen to reduce time and costs and to limit engine wearing while obtaining an acceptable degree of accuracy. It is also proved that the tests to assess the improvements obtained with design changes could not have been completed in the steady-state mode using a single engine because of the short life cycle of racing engines due to wearing, which would have altered the comparisons. Test results show a 16% and 33% rise in torque and power for the racing configuration, reaching the state of the art of the best performing engines in the Italian Supersport racing class. Ó 2009 Elsevier Ltd. All rights reserved.
1. Introduction The search for specific power gain of high speed internal combustion engines has always been a critical issue in F1 and MotoGP prototypes races. This search presents a long history also in the field of the stock production engines modifications both for racing purposes and for research and technological development. It is well known that under the pressure of the increasing fossil fuel costs almost all automotive industries involved in the production of passenger car engines are reviewing their production in the direction of ‘‘downsizing” philosophy, which necessarily requires increasing engine specific power. Many detailed descriptions of the modifications that are suggested to improve specific power of stock engines are available in the literature. The results obtained obviously depend on the complexity of the modifications allowed. Thus, as examples, in the early sixties Ferro and Martegani [1] almost doubled the specific power of the small passenger car 479 cm3 FIAT 110 two cylinders carburetted engine by a pioneering gasoline direct injection application, while more recently O’Blenes and Bothwell [2] transformed the 105 kW GM Ecotec 2198 cm3 into a 744 kW sport compact drag racing methanol fuelled engine equipped with an innovative turbocharger capable of producing 0.27 MPa of boost, and Engel et al. [3] obtained 320 kW by the 5.670 cm3 GM LS1 * Corresponding author. Tel.: +39 049 8276746; fax: +39 049 8276788. E-mail address: [email protected] (M. Masi). 0306-2619/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2009.09.033
V8 engine adapted for the ASA Stock Car racing series following a strategy of cost savings. Modifications to the intake and exhaust systems aimed at improving volumetric efficiency are very effective to gain specific power. Many standard books in the literature report theory and experimental data about aerodynamics of engine head manifolds, (see, e.g., Heywood [4] and Stone [5]). Since the 1960s the research of Kastner et al. [6] has well established the effects of poppet valve characteristics on engine performance. Woods and Khan [7] and Annand and Roe [8], among others, are classical works dealing with the head manifolds discharge coefficients. A big effort was done to study acoustic waves effects into intake and exhaust systems in order to improve global performance of internal combustion engines. Winterbone and co-workers reviewed a century of researches on the analysis and design of internal combustion engine breathing systems [9]. A review on design and simulation of high-performance engines is due to Blair [10]. Many insight about the methodological approach to design and performance estimate of motorbike high speed racing engines are given by Boretti [11] and related references, and recently by Mattarelli [12]. While theory and practice on engine testing are intrinsically interconnected to the advances in internal combustion engines technology (see, e.g., the classical text by Judge in the 1950s [13] or the more recent by Plint and Martyr [14]), only few references exist about testing of high speed racing engines. Among these, Harrison and Dunkey [15] study 600 kW V10 F1 engines of the early 2000s to develop a decoupled hybrid method
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Nomenclature reference area AR BBDC ABDC before and after bottom dead centre BTDC ATDC before and after top dead centre CA crankshaft angle discharge flow coefficients Cd, Cf EO, EC exhaust valve open, exhaust valve close IO, IC inlet valve open, Inlet valve close, moment of inertia at the crankshaft of all the engine Ieng moving components moment of inertia at the crankshaft of the whole driveIdl line moment of inertia at the crankshaft of applied load IL um mean piston speed um° mean piston speed at maximum power (from steadystate tests on the modified engine clutched on 4th gear) bmep brake mean effective pressure
calculating wave dynamics in the intake system of single cylinder racing engines. They perform steady-state tests on single cylinders similar to those of the real racing engine, as some F1 manufacturers do in order to reduce the costs involved in engine testing. In spite of the advantages of this testing criteria, the authors clearly report the difficulty in reproducing the V10 engine performance by means of single cylinder tests. The reason of this are the unsteady cyclic nature of the intake flow, the friction losses per cylinder, which are higher in the single cylinder engine due to extra mechanical systems required to balance the engine, and the absence of exhaust tuning effects due to the common exhaust of the cylinders in the same bank. To account for a more realistic contribution of the exhaust gases blow-down, Blair suggests to test the complete racing engine by the slow speed ramp testing mode technique [16]. This old technique consists in measuring the main engine performance parameters at fixed throttle position while engine rotational speed is increased within a range in a relatively long time (10–12 s). It was adopted among the others by Weslake in the 1950s to develop race engine designs and is widely used today by technicians involved on high speed engine tuning despite of the lower accuracy associated with the testing mode. In spite of these theoretical and methodological references, the availability of measured data about very high specific power internal combustion engines for racing motorbikes is not so high. In his renowned book [17], Blair provides very useful empirical guidelines for designers and some interesting experimental data about the performance of high speed engines, including those about the V-twin 955 cm3 FIM Superbike Ducati engine from the work by Boretti et al. [18]. This work presents the results of an experimental analysis used to validate numerical calculations aiming at predicting engine performance. The experimental data are shown as nondimensional ratios with respect to maximum torque and power, but engine mean effective pressure is known to be higher than 1.4 MPa. The same scaling technique of the experimental data is used in Cantore and Mattarelli [19] about the performance of a 2003 prototype MotoGP three-cylinder inline engine. A methodology for a parametric design based on similarity rules is developed, and the performance curves of 1.3 MPa mean effective pressure MotoGP engines, obtained by 1D gas-dynamics numerical simulations of different engine layouts, are presented. These curves include a four-cylinder inline configuration, the same of the engine that is studied here. This paper presents an experimental study [20] of a high speed stock motorbike engine [21], which was modified in compliance
bmep°
dLmep p0, T0 pT R T TL Tdl W WOT
brake mean effective pressure at maximum power (from steady-state tests on the modified engine clutched on 4th gear) driveline absorption in terms of mean effective pressure upstream throttle stagnation pressure and temperature throttle pressure gas constant engine torque at crankshaft load torque absorption driveline torque absorption mass flow rate wide open throttle
Greeks
c x
specific heat ratio rotational engine speed
with the 2007 FIM Road Racing Supersport Italian Championship (CIV) [22] to gain specific power. After a synthetic description of the engine modifications, the attention manly focuses on: – The experimental results about the aerodynamic performance of the modified intake and exhaust ducts inside the engine head, involving discharge coefficients measurements on a steady-flow test rig obtained by modifying an industrial fan test rig [23]. – The reliability of the slow speed ramp test performed on the chassis dynamometer as a suitable testing technique for racing engines. – Brake torque, power and specific fuel consumption measurements both at wide open throttle and at part load. – Power losses measurement due to driveline end differences between the engaged gears. Finally the effects of some valve timing parameters on performance are also evaluated. Presented results show torque and power increases for the modified engine of 16% and 33%, respectively. The bike competed in the final part of the 2007 Supersport Italian Championship, scoring the 10th position in the overall standings, the best results obtained being a 4th and a 6th position. 2. The engine The engine powering the Kawasaki ZX-6R/07 motorbike [21] is modified to comply with the standards imposed by the FIM Road Racing Supersport Italian Championship (CIV) [22], that is the racing class for medium volume motorbikes derived from stock production. Table 1 reports the engine technical specifications that are essential for the experimental work developed in the paper. The modifications to the original engine, which are briefly summarised in the following, are all aimed at improving output power within the limitations imposed by the regulations. Combustion chamber shape and volume are modified by altering cylinder head in order to raise the compression ratio (special gasoline is adopted to avoid detonation). Fuel injection system ECU is substituted, and ignition advance and injection timing and duration maps are experimentally tuned as functions of regime and load. Camshafts are changed, in particular cam profiles are shaped to better perform at the higher rotational speeds while valve lift is
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Table 1 Engine technical specifications. Engine model type architecture
Kawasaki ZX6R-07 4 stroke spark ignition inline 4 cylinders
Unmodified stock engine data Swept volume Stroke/bore Combustion chambers Firing order Intake valves Exhaust valves
Valve train Injection system
599 cc 0.634 4-valves pent-roof type 1–2-4–3 from left to right Outer seat diameter 27.5 mm; lift 8.3 mm; lash 0.2 mm Outer seat diameter 23.1 mm; lift 7.5; lash 0.25 mm.
Cooling
Direct drive, flat valve lifters with DOHC chain driven, helicoidal springs Multi point port fuel electronic injection; 2 racks of 8 holes injectors (rack Keihin in the airbox, passive up to 7000 rpm, rack Mitsubishi in the intake ports gradually shut off over 7000 rpm) Liquid cooled forced circulation by radial pump crankshaft geared
Engine parameter/component
Value/intervention
Modifications at the stock engine Compression ratio Fuel Equivalence ratio ECU Piston speed Camshaft Valve springs Valve timing Exhaust system Cooling system Lubricant
14.5 WLADOIL unleaded racing fuel (M.O.N. 89.1 ± 0.4) Ref.: WLADOGAS BB1 Cod. 22544600 1.11 MoTeC m800 22.66 m/s Changed Changed IO: 48° BTDC; IC: 82° ABDC (310° total crank angle opening) EO: 71° BBDC; EC: 39° ATDC (290° total crank angle opening) Substituted with the Akrapovic S-K6EFT5T-ACT kit Supersport 2007 Substitution of the original heat exchanger with a unit able to dissipate a maximum thermal load of about 80 kW SAE 5 W-40 oil
kept constant. Valve springs are changed as well as a necessary consequence. No changes in mechanical components are allowed to reduce mechanical friction, so the only way to improve organic efficiency, apart from surface finishing, is to use a better lubricant. The original coolant heat exchanger is substituted with another unit that is placed in the same location but has a surface that is three times as much. The two following subsections deal with the air breathing system components and the modifications applied to them in order to increase the engine volumetric efficiency. 2.1. Intake and exhaust systems The upper part of Fig. 1 shows the path of the fresh charge inside the air intake system that supplies air to all the cylinders. For sake of simplicity this system is subdivided into three main subsystems: the secondary pipe, the air filter box, and the primary manifolds. The secondary pipe (‘‘airscoop” in Fig. 1) features a large diffuser that is split into two rectangular ducts crossing the steering zone and benefits from ram charging at high speed due to its inlet on bike front. The air filter box (‘‘airbox” in Fig. 1) is a 7.5 l plenum between the secondary pipe and cylinder intake runners. Each of the four independent primary manifolds is made of a short bellmouthed duct (‘‘intake trumpet” in the lower part of Fig. 1), a throttle body equipped with two throttle valves in one throttle bore (a secondary ECU controlled valve is placed upstream the primary one which is used for load control), and the two port passages inside the engine head. Only slight modifications were made to the intake system layout upstream the ports in the engine head, namely:
(1) the bell-mouthed ducts were shortened to rise their tuning frequency, (2) the secondary throttle valves were kept at their maximum opening, as their role is not significant at the rotational speed range of interest, (3) the original air filter was substituted with a 3 mm sponge filter to reduce the pressure loss (although a minimal benefit was achieved). The exhaust system beyond the engine head is completely substituted according to FIM regulations [22], maintaining nothing but the location and number of the original silencers. The exhaust piping downstream engine head is assembled from steel tube components and presents the classical ‘‘4-into-2-into-1” configuration (see Fig. 2). The two intermediate secondary branches collect the primary exhausts of cylinders 1–2 and those of cylinders 3–4, respectively, and discharge the exhaust gases through an Y-junction in the common tailpipe that ends in the silencer. This layout enhances power output at the high speed range of engines with 1-2-4-3 firing order [24]. In addition, Fig. 2 shows two ‘‘interference” pipes (16 mm diameter, and 35 mm and 40 mm length, respectively) connecting the primary manifolds before their Y-junction with the secondary branches. Table 2 reports the main dimensions of the engine air breathing system. The silencer (see the lower part of Fig. 2) is of absorption-type, because of its low exhaust counterpressure [25]. It is made of a 1 mm external steel jacket (section of 13341 mm2) filled with an absorbent lining that covers a 380 mm steel pipe (diameter 54 mm, thickness 1 mm) with a perforated central part (length
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Fig. 1. Intake system (above) and details of the external air section (below).
Fig. 2. Exhaust system: exhaust line (upper-left), details of the interference pipes (upper-right), silencer (lower-left) and its simplified scheme (lower-right).
273 mm). Holes diameter and spacing are about 2 and 1.5 mm, respectively.
isentropic steady-flow discharge coefficient Cd, defined for non chocked flows as:
2.2. The head and the discharge coefficients measurement
" 1=c ( ðc1Þ=c #)1=2 WðRT 0 Þ1=2 pT 2c p Cd ¼ 1 T c1 AR p0 p0 p0
The shape of intake and exhaust ports through cylinder head are modified and piping internal surface has been polished to improve its aerodynamic performance. This can be expressed in terms of the
where W is the actual mass flow rate, p0 and T0 are the upstream pressure and temperature stagnation condition, pT is the static pressure just downstream of flow restriction, g is the specific heat ratio,
ð1Þ
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M. Masi et al. / Applied Energy 87 (2010) 1641–1650 Table 2 Main dimensions of external air section and exhaust system. System
Subsystem
Intake
Secondary pipe (airscoop)
Ram intake Steering zone
Air filter box (airbox)
Zone A
Zone B Zone C
Exhaust
Primary exhaust manifolds
Secondary branches Tailpipe
Length (mm)
Equivalent diameter (mm)
Volume [l]
0 280 0 140 0 95 215 – 0 50
67.8 98.37 69.56 (2) 87.4 (2) 123.62 123.62 190 – 105.25 (2) 108.82 (2)
–
28.7 30 38 38 40 40 43 43 54 54
– – – – – – – – – –
Cyl. 1
Cyl. 2
Cyl. 3
Cyl. 4
0 10 145 445 0 180 0 160 250 1070
0 10 150 475
0 10 155 465 0 180
0 10 140 455
and AR is a reference section that can be defined in several different ways (see Heywood [4]). The geometry of the modified ports (see Fig. 3 for the main dimensions of the equivalent rectified geometry) is acquired by applying a laser scanning technique to a rubber ‘core’ obtained by filling the empty space within the ports. Surface data are then handled with the RapidformÒ 3D Scanning Software (see Fig. 4) to get area measurements at various cross-sections. The aerodynamic performance of the modified intake and exhaust ports is measured on an experimental apparatus derived from a type A industrial fan test rig that is built according to the UNI 10531 Italian Standard [26], which is equivalent to the ISO 5801 Standard. The main scheme of the original test rig, described in details by Martegani et al. [23], is made of an auxiliary fan that blows into a test chamber on the suction side of the fan to be tested. In this type of layouts, an orifice plate is
– 3.6
3.0 0.9
provided by the standard for flow rate measurements. The facility has been modified by substituting the auxiliary fan with a variable speed fan that generates a negative relative constant pressure inside the plenum chamber. A one-cylinder slice of the engine head replaces the fan to be tested. In this way, a constant pressure drop is kept between ambient pressure and plenum chamber pressure across the opened ports, so that the air flow rate can be measured as a function of valve lift only. Dry and wet-bulb ambient temperature and the temperature inside the plenum chamber are measured using mercury thermometers. Barometric pressure is measured using a mercury barometer, whereas a water manometer is used to measure the pressure in the plenum chamber. The orifice plate differential pressure, that is needed for flow rate determination, is measured by a differential water manometer. The accuracy estimated by the Kline–McClintock criterion [27] is less than 1.3% for all the presented discharge coefficient measures. Both direct and reverse flow discharge coefficients are experimentally obtained for different values of the pressure ratio across the ports. Fig. 5 illustrates a sample of the discharge coefficients values of intake and exhaust ports at a fixed pressure ratio of 1.035. In particular, Fig. 5a shows the values of the discharge coefficients Cd that is normalised with respect to a fixed reference section AR (the area of the valve inner seat) as proposed by Woods and Khan [7]. Fig. 5b shows the values of the discharge coefficients Cf that is obtained adopting the ‘‘valve curtain area”, which is a linear function of the valve lift, as reference section AR [8].
duct area / piston area
0.60
3. Engine test rig 0.50 0.40
exhaust intake
C
B
A
0.30 0.20 0.10 0.00 0
unchanged until 1sth Y junction
engine head
40
80
120
160
200
240
distance from valve [mm] Fig. 3. Primary intake (above) and exhaust manifolds (below): sketch and main dimensions.
The engine is tested in an acoustic-insulated cell (Fig. 6) equipped with an intake air fan, which supplies air for engine cooling and breathing (flow rate: 0.37–1.32 m3/s; total pressure: 1176– 1715 Pa), and an exhaust extractor, which removes combustion gases from the room. Inside the cell, a Dynojet Mod. 250i chassis dynamometer (flywheel inertia allows 373 kW maximum power absorption) is coupled with the Jacobs Mod.12JC30 air-cooled eddy current dynamometer and a 1162.5 Nm torque cell. The instrumentation also include sensors for ambient air temperature, pressure and humidity. The results of the tests (which are governed by an elec-
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Fig. 4. Picture and digital acquisition of intake (left) and exhaust (right) pipes inside the engine head.
a
0.8 0.7 0.6
Cd
0.5 0.4 0.3 Cd intake Cd intake reverse Cd exhaust Cd exhaust reverse
0.2 0.1 0.0 0
0.05
0.1
0.15
0.2
0.25
0.3
0.35
0.4
lift / inner seat diameter
b
0.8 Cf intake Cf intake reverse Cf exhaust Cf exhaust reverse
0.75 0.7 0.65
Cf
0.6 0.55 0.5 0.45 Fig. 6. The engine test cell.
0.4 0.35 0.3 0
0.05
0.1
0.15
0.2
0.25
0.3
0.35
lift / outer seat diameter
The test rig, both in transient and in steady-state mode, measures wheel performance. This means that the crankshaft engine torque T can be calculated only if all the terms on the right hand side of the following equation are known:
Fig. 5. Direct and reverse isentropic discharge flow coefficients for the intake and exhaust ports defined by different reference sections: (a) the area of the valve inner seat and (b) the ‘‘valve curtain area”.
T ¼ T dl þ T L þ ðIL þ Idl þ Ieng Þ
tronic module) are instantly corrected following the ISO 3046-1 Standard [28] before real time display and storage on a PC. Both steady-state and transient measurements can be performed in this facility. Engine speed measurement is derived from the ECU pulse sent to cylinder 1 ignition coil and its resolution is 1 rpm. According to the manufacturer, the accuracy on other quantities related to the chassis dynamometer are: timing accuracy of 1 ms; drum speed accuracy of 1/100 mi/h; rpm accuracy 1/10 rpm. Torque accuracy can be estimated as 0.2% of the torque full scale reading, that is ±2 Nm, in the steady-state tests carried out using the electric dynamometer. Fuel consumption is obtained by a weight measure. The fuel tank is arranged on a load cell with a resolution of 1 g and a maximum linearity error of 2 g. The acquisitions are completed by the air–fuel ratio measurement, derived from the signal of the lambda sensor.
where x is the rotational engine speed, TL is the applied load, IL, Idl and Ieng are the moments of inertia of the applied load, driveline and engine, respectively. Thus, in order to obtain the brake performance curves, it is necessary to determine the driveline losses Tdl induced by the gearbox and the transmission (this is necessary for the simpler steady-state tests (dx/dt = 0) as well). To this end, a ‘‘coast down” test is carried out (driveline inertia Idl must be known or estimated), in which the engine of the geared bike is first brought to maximum rotational speed, then the clutch is disconnected manually (T = TL = Ieng = 0) and torque is acquired during the free deceleration of the complete driveline led by the inertia of the chassis dynamometer flywheel. The description of performance curves at a fixed throttle position for a specified engine setup by traditional steady-state tests requires at least 20 acquisitions at different rotational speeds, so the engine works for no less than 30 min roughly. As a consequence, a single racing engine, which is optimised for a few hours
dx ; dt
ð2Þ
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4.5%
4.5% dLmep: experiments dLmep: data fitting dLmep*um: experiments dLmep*um: cubic regr. data fitting
dLmep / bmep°
3.5%
4.0%
[dLmep x um] / [um° x bmep°]
4.0%
3.5%
3.0%
3.0%
2.5%
2.5%
2.0%
2.0%
1.5%
1.5%
1.0%
1.0%
0.5%
0.5%
0.0% 20%
0.0% 40%
60%
80%
100%
um / um° Fig. 7. 4th Gear driveline losses.
a
105%
110% 100%
95%
90%
85%
80%
75%
70%
65% 55%
60%
bmep: sweep test bmep: steady-state test
50% 40% 30%
bmep*um: sweep test bmep*um: steady-state test
40%
50%
60%
70%
80%
45%
[bmep x um] / [bmep°x um°]
bmep / bmep°
4th gear tests
35% 90% 100% 110%
110% 100%
4th gear - SST
90%
bmep / bmep°
operation, cannot bear several full load tests at many rotational speeds. Other researchers (see, e.g., Blair et al. [16]) suggest to overcome this limitation by changing the testing approach for racing engines and carrying out an experimental campaign mainly based on short transient runs, i.e. slow speed ramp tests briefly called ‘‘sweep tests”. The complete performance curves (except
80% 70% 60% engine 1
50%
engine 2
40% 30%
40%
50%
60%
70%
80%
90%
100% 110%
um / um° Fig. 9. Steady state test mean effective pressure measurements (4th gear) for new engine (engine 1) and worn engine (engine 2).
for the data about specific fuel consumption) are obtained from a 10 s-to-12 s sweep test. Regardless of measurement uncertainty, which is due to the estimation of Ieng and Idl in Eq. (2), this testing mode is certainly the least stressing for the engine. The opportunity to perform the major part of the tests in sweep mode, even if its reliability is lower than the more precise braking mode, is probably the only way to test different setups on such an engine at the same wear level in all tests (see the discussion about Fig. 9). To evaluate the repeatability of the measurements, all the slow speed ramp tests are repeated at least three times. Because of the reduced number of test repetitions, accuracy is determined by means of the Student criterion with a 90% confidence, as established in the European Standards [29]. Repeatability errors in torque measurements for the engine revving range shown in all the presented figures are under 8% in the worst case and the mean error is under 2% in almost all tests (the worst case mean error value is below 4%, this limit being claimed as typical when testing highperformance engines [16]; see the discussion below about Fig. 8a). Most of the tests are performed with the engine clutched on 4th gear, to guarantee operating stability in almost all the rotational speed range that is useful during a competition. 4. Experimental results
um / um° 105%
110%
bmep / bmep°
4th gear tests
100%
95%
90%
85%
80%
75%
70%
65%
60%
55%
50%
bmep: modified engine bmep: stock engine 45% bmep*um: modified engine bmep*um: st ock engine
40% 30%
40%
50%
60%
70%
80%
90%
[bmep x um] / [bmep°x um°]
b
35% 100% 110%
um / um° Fig. 8. Brake mean effective pressure and power per piston area measurements (4th gear): (a) comparison between sweep test and steady-state test at WOT and (b) comparison between the stock and the modified engine.
This section presents the experimental results about the engine performances in its racing setup. Although torque and rotational speed are measured directly in the test rig, all the experimental results in the present paper are shown in terms of brake mean effective pressure (bmep) and piston speed (um), and are made nondimensional using the factors bmep° (1.38 MPa) and um° (20.6 m/ s), which are the brake mean effective pressure and the piston speed, respectively, measured by a steady-state mode test in which the engine operates in its racing setup at maximum power clutched on 4th gear. Fig. 7 shows driveline losses with the 4th gear engaged in terms of mean effective pressure (dLmep) and of power per piston area (dLmep um). Both non-dimensional curves are obtained from a cubic regression fitting of mean effective pressure measured data. As stated in the previous section, some test are carried out in steady-state mode to compare the results with sweep-mode measurements and to estimate the accuracy level of the latter. The curves in Fig. 8a show the wide open throttle (WOT) performance for the engine in its racing setup as obtained by the sweep-mode tests. Brake mean effective pressure and power per piston area values at WOT condition as measured in steady-state mode tests are
M. Masi et al. / Applied Energy 87 (2010) 1641–1650
plotted as dots in the diagram as well. The error bars of the sweep test curve show the estimate of the random instrumentation errors only (see Plint and Martyr [14]). It is clear that the agreement between the results obtained with the two test modes is not complete, especially at low rotational speeds. While differences at low regimes are strongly affected by engine instability, sweep measurements also suffer by the rough evaluation of the overall rotating part inertia Ieng and Idl in Eq. (2) (it was not possible to perform massive tests because of motorbike availability and the need of preserving against wear an engine ready for competitions). Moreover, exhaust pipe wall and exhaust gas temperatures, and then wave propagation velocities, are not identical in the two test modes [30]. However, although the error band is quite large, the diagram shows that the results of the sweep test mode are well inside the band of accuracy of the traditional steady-state test for almost all the rotational speed range of practical interest. To enforce the advantages deriving from the sweep test technique, the performance of two identical engines were compared in Fig. 9. The former, numbered as engine 1, is the one used in this work, the latter, numbered as engine 2, worked in competitions for about 1000 km, undergoing a number of working cycles comparable with those of an extensive test bed experimentation based on steady-state tests. Despite of the amplitude of the uncertainty, the tests on these engines performed in the steady-state mode clearly quantify the mean incidence of wearing on torque as 2% of the maximum brake mean effective pressure, with peaks that are greater than 5% at some revving speeds. Fig. 9 also shows clearly that the reliability and the duration of the modified engine are much worse. This drawback is however emphasized by the technical regulations [22] that do not allow the substitution of many mechanical components which would require different designs and materials to keep up with the increased performance. In Fig. 8b the brake mean effective pressure and power per piston area curves of the modified engine described in Section 2 are compared with the corresponding performances data measured for the stock engine. The sweep-mode tests highlight a consistent performance gain for the modified engine at all the regimes above 45% of um°. Moreover, the increment in maximum brake mean effective pressure for the modified engine is about 16% (from 1.22 to 1.42 MPa), while maximum power per piston area shows an improvement of more than 33%. The comparison between the two engine configurations shows that the bmep curve maintains approximately the same shape, but, as expected, the trends are sharpened with higher peaks and lower drops for the modified configuration. The most substantial modifications to the air intake system and the total replacement of the exhaust system result in an improvement of the volumetric efficiency which occurs at the higher rotational speeds in particular. Performance is therefore enhanced at high regimes, whereas it drops significantly at lower regimes. This may be viewed as a potential drawback for stock engines, but racing engines rarely operate below 60% of their maximum power rotational speed. Fig. 10 shows the specific fuel consumption evaluated in a steady-state test in which the time required to use up to 100 g of gasoline is recorded, the uncertainty of this indirect measure being estimated using the Kline–McClintock criterion [27] as suggested by the standard [29]. The curve shows some peaks (the highest occurs at about 50% of um°) at the regimes featuring volumetric efficiency drops due to the increasing in pumping work. Note, however, that the magnitude of these peaks is not very significant if the scale on the ordinate axis is considered. A full comparison with similar high technology racing engines is difficult, due to the little availability of experimental data. A paper by Cantore and Mattarelli [19] reports some non-dimensional experimental data about torque and specific fuel consumption for a three-cylinder MotoGP year 2003 prototype. While the
340 Specific Fuel Consumption
330 320 310
[g/kWh]
1648
300 290 280 270 260 250 240 30%
40%
50%
60%
70%
80%
90% 100% 110%
um / um° Fig. 10. Specific fuel consumption of the engine.
three-cylinder MotoGP engine offers appreciable differences in specific power, and more important, in power per piston area (they are about 8% and 14% greater than those of the engine tested here), the performance curves exhibit some agreement about the location and the relative amplitudes of peaks. This is because the air breathing system layout, except for the telescopic taper device that can vary continuously intake pipe length, is mostly independent of the technological level within the very high speed engines category. Note that the brake mean effective pressure of the engine tested here is about the same of those tested in [18] and [19]. Thus the results presented for this SuperSport racing motorbike engine have a general validity for very high speed motorbike engines including the MotoGP ones, despite to their higher technology level (comparable to that of F1 engines [19,12]). Fig. 11 illustrates the WOT brake mean effective pressure and the increment in transmission losses (in terms of mean effective pressure) when the 6th gear of the driveline is engaged instead of the 4th. The average of the difference in mean effective pressure absorption at all the rotational speed is 1.8% of bmep°. Fig. 12 illustrates brake mean effective pressure performance at different loads for the driveline with the 6th gear engaged. The dotted line in the diagram is the 100% full throttle (WOT) brake mean effective pressure curve obtained from other tests carried out with the driveline 4th gear engaged. The addition of the average of the difference in mean effective pressure absorption between 4th and 6th gear (Fig. 11) confirms the previous rough estimate of this value and indirectly confirms once again the acceptable degree of repeatability for the measures. As expected, part load curves show the performance falling down because of the deterioration of the volumetric efficiency, which is strongly affected by the throttling imposed by load control. Please note the trend at the medium to low rotational speed where the 50% full load torque curve exhibits performance data that are better than those of the 75% torque curve. This behaviour is often featured by engines the valve timing of which is optimised for high rotational speeds (e.g., racing engines), where a strong fresh charge backflow in the intake manifolds occurs at the beginning of the compression stroke for low regimes. In such engines the throttling by the load control system at low regimes acts like a barrier to backflows, and it is not unusual that further throttling increases brake mean effective pressure. This is the main reason of the secondary throttle valve in almost all the stock engines, as confirmed by Wakimura and Nitta [31]. On the other hand, the secondary throttle has been locked wide open in these tests because of the high operation regimes of the engine. Finally, Fig. 13 shows non-dimensional brake mean effective pressure and power per piston area curves obtained with an ad-
M. Masi et al. / Applied Energy 87 (2010) 1641–1650
110%
4.0%
6th gear 4th gear
100%
3.5%
90%
3.0%
80%
2.5%
70%
2.0%
60%
1.5%
50%
1.0%
40%
0.5%
30% 30%
40%
50%
60%
70%
80%
difference / bmep°
bmep / bmep°
difference
0.0% 90% 100% 110%
um / um° Fig. 11. Brake mean effective pressure on 6th and 4th gear and difference in mep absorption between gears (WOT).
110% 100%
bmep / bmep°
90% 80% 70%
25% full throttle
60%
50% full throttle
50%
75% full throttle 100% full throttle
40% 30% 30%
100% full throttle 4th gear corrected to 6th
40%
50%
60%
70%
80%
90% 100% 110%
um / um°
110%
105%
100%
95%
90%
85%
80%
75%
70%
65% 55%
bmep MOP 104°
60%
bmep MOP 107°
50%
bmep*um MOP 104°
45%
50%
60%
70%
80%
90%
100%
5. Conclusions An experimental investigation on the performance of a motorbike high speed racing engine was presented. The modifications to the stock engine and the geometry of the complete air breathing system, in compliance with the FIM Supersport race class regulations, were illustrated and experimentally tested. Results show a gain of about 16% in maximum torque and 33% in maximum power for the modified engine. A comparison with the performance of an identical engine having expired its short racing life cycle pointed out that massive steady-state runs at the test bed could not be performed because of performance degradation due to wearing, which is of the same order of magnitude of the performance improvement generated by specific design changes. Thus, engine performance at both wide open throttle and part loads, intake valves timing effect, driveline and gearing influence were analysed by slow speed ramp testing mode on a chassis dynamometer test bench. The worst case mean error value on torque measurements is below 4%. The comparison with steady-state measurements and the error analysis revealed an acceptable degree of accuracy for the slow speed ramp testing mode even for a low number of test repetitions, and, therefore, suggests this testing mode is suited to high speed racing engine in order to save wearing, costs and time. Finally, the indications provided here can be shared among different designs by using similarity rules, and the presented measures both for discharge coefficients and gross performance offer useful data for one dimensional gas dynamics engine simulations.
The authors gratefully acknowledge Giorgio Mozzo for the great tuning engines experience he shared in this work, and Andrea Mozzo for the great help in the dynamometer tests.
bmep*um MOP 107°
40% 40%
imum gain in brake mean effective pressure (and power per piston area) of about 4% in the range from 48% to 55% of um°, with a little decrease at the maximum power per piston area regime (however, the difference of about 1% is comparable with the local measures uncertainty). This result suggests that additional modifications to the camshafts may be beneficial.
Acknowledgments
[bmep x um] / [bmep° x um°]
bmep / bmep°
Fig. 12. Brake mean effective pressure at different loads (6th gear).
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35% 110%
um / um° Fig. 13. Brake mean effective pressure and power per piston area curves for MOP at 104° ATDC and 107° ATDC.
vance of intake valve mean opening period (MOP) angle from 107° after top dead centre (ATDC) to 104° ATDC. The advance, which anticipates intake valves closure, is made by an azimuth shift of the intake camshaft. Brake mean effective pressure is improved in almost all the medium/high range of regimes, obtaining a max-
References [1] Ferro A, Martegani, AD. Ricerche sulla alimentazione ad iniezione diretta di benzina in un motore automobilistico di piccola cilindrata, Ingegneria Meccanica, anno XIII, no.5; 1964. [2] O’Blenes RD, Bothwell SG. Ecotec 2.0 liter race engine. SAE paper 2002-013354; 2002. [3] Engel D, McPhail M, Taylor D. Putting the stock back in stock car racing: developing the GM LS1 V8 for short-track competition. SAE paper 2002-013356; 2002. [4] Heywood JB. Internal combustion engine fundamentals. New York: McGrawHill; 1988. [5] Stone R. Introduction to internal combustion engines. 3rd ed. London: MacMillan Press Ltd.; 1999. [6] Kastner LS, Williams TJ, White JB. Poppet Inlet valve characteristics and their influence on the induction process. Proc I Mech E 1963;178:955–75. [7] Woods WA, Khan SR. An experimental study of flow through poppet valves. Proc I Mech E 1965-1966;18(32). [8] Annand WJD, Roe GE. Gas flow in the internal combustion engine. Newfoundland, NJ: Haessner Publishing; 1974. [9] Winterbone DE, Pearson RJ. Design techniques for engine manifolds: wave action methods for IC engines. London: Professional Engineering Publishing Ltd.; 1999. [10] Blair GP. Design and simulation of engines: a century of progress. SAE paper 1999-01-3346; 1999. [11] Boretti AA. Parametric design of FIM WGP engines. SAE paper 2002-01-3317; 2002. [12] Mattarelli E, Moto GP. 2007: criteria for engine optimisation. J Eng Gas Turb Power 2008 [130/015001-1].
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[13] Judge AW. The testing of high speed internal combustion engines. Chapman & Hall; 1954. [14] Plint M, Martyr A. Engine testing – theory and practice. 1st ed. ButterworthHeinemann Ltd.; 1995. [15] Harrison MF, Dunkley A. The acoustic of racing engine intake systems gas dynamics in racing engines. J Sound Vib 2004;271:959–84. [16] Blair GP, Mackey DO, Ashe MC, Chatfield GF. Exhaust tuning on a four-stroke engine; experimentation and simulation. SAE paper 2001-01-1797/4218; 2001. [17] Blair GP. Design and simulation of four-stroke engines. SAE, R-186, ISBN 07680-0440-3; 1999. [18] Boretti AA, Cantore G, Matterelli E, Preziosi F. Experimental and computational analysis of a high performance motorcycle engine. SAE paper 962526; 1996. [19] Cantore G, Matterelli E. Similarity rules and parametric design of four-stroke motogp engines. SAE paper 2004-01-3560; 2004. [20] Mozzo A. Analisi Numerica e Sperimentale di un Motore Motociclistico da Competizione, (numerical and experimental analyses of a motorbike racing engine), MSc Thesis, University of Padova – Dept. Mech. Eng., Padova, Italy; 2007 [in Italian]. [21] Kawasaki Heavy Industries, Ltd. Kawasaki Ninja ZX6R-07 – manuale di assistenza della motocicletta (Kawasaki Ninja ZX6R-07 – motorbike technical guide manual); 2007 [in Italian].
[22] Fédération Internationale de Motocyclisme (FIM). Road racing superbike and supersport world championships and superstock cup regulations. 1st ed.; 2007. [23] Martegani AD, .Lazzaretto A, Taffurelli A. A user’s guide to the UNI 10531 standard on performance testing of industrial fans. Document ISO/TC117 N306 Annex 3, 4 December; 2003. [24] Nakayasu T, Yamada H, Suda T, Iwase N, Kyo T. Intake and exhaust systems equipped with a variable valve control device for enhancing of engine power. SAE paper 2001-01-0247; 2001. [25] Smith PH, Morrison JC. Scientific design of intake and exhaust systems. 3rd ed. Bentley Publishers; 1971. [26] UNI 10531. Ventilatori industriali – metodi di prova e condizioni di accettazione (Industrial fans – test methods and conditions for the acceptance), Milan; 1995 [In Italian]. [27] Kline SJ, McClintock FA. Describing uncertainties in single sample experiments. Mech Eng 1953. [28] ISO 3046-1. Reciprocating internal combustion engines – performance; 2002. [29] ENV 13005. Guide to the expression of uncertainty in measurement; 1999. [30] Blair GP, Mackey DO, Ashe MC, Chatfield GF. On improving the accuracy of the simulation of reciprocating engines. In: ImechE conference, C587/001/2000; 2000. [31] Wakimura M, Nitta Y. Development of a fuel injection system for highperformance motorcycles. SAE paper 2002-32-1824; 2002.
Contents lists available at ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Experimental analysis of a motorbike high speed racing engine Massimo Masi *, Andrea Toffolo, Marco Antonello Department of Mechanical Engineering, University of Padova, Via Venezia 1, 35131 Padova, Italy
a r t i c l e
i n f o
Article history: Received 27 April 2009 Received in revised form 25 September 2009 Accepted 25 September 2009 Available online 25 October 2009 Keywords: High speed engine testing Sweep test reliability Discharge flow coefficients
a b s t r a c t Power gain is the main objective in any motorbike competition. Despite of the wide literature on theoretical and experimental methods for increasing engine power, there is a general lack of data about tests on racing engine performance due to the obvious manufacturers’ reluctance to spread information, especially for recent high technological level applications. This paper, instead, presents all the main results of the experimental tests conducted on a motorbike engine both in the original stock arrangement and in a modified configuration proposed in compliance with the Technical Regulations of the 2007 FIM Road Racing Supersport Italian Championship (CIV). Traditional testing techniques (steady-flow discharge coefficients measurements and chassis dynamometer tests performed in the slow speed ramp mode) are chosen to reduce time and costs and to limit engine wearing while obtaining an acceptable degree of accuracy. It is also proved that the tests to assess the improvements obtained with design changes could not have been completed in the steady-state mode using a single engine because of the short life cycle of racing engines due to wearing, which would have altered the comparisons. Test results show a 16% and 33% rise in torque and power for the racing configuration, reaching the state of the art of the best performing engines in the Italian Supersport racing class. Ó 2009 Elsevier Ltd. All rights reserved.
1. Introduction The search for specific power gain of high speed internal combustion engines has always been a critical issue in F1 and MotoGP prototypes races. This search presents a long history also in the field of the stock production engines modifications both for racing purposes and for research and technological development. It is well known that under the pressure of the increasing fossil fuel costs almost all automotive industries involved in the production of passenger car engines are reviewing their production in the direction of ‘‘downsizing” philosophy, which necessarily requires increasing engine specific power. Many detailed descriptions of the modifications that are suggested to improve specific power of stock engines are available in the literature. The results obtained obviously depend on the complexity of the modifications allowed. Thus, as examples, in the early sixties Ferro and Martegani [1] almost doubled the specific power of the small passenger car 479 cm3 FIAT 110 two cylinders carburetted engine by a pioneering gasoline direct injection application, while more recently O’Blenes and Bothwell [2] transformed the 105 kW GM Ecotec 2198 cm3 into a 744 kW sport compact drag racing methanol fuelled engine equipped with an innovative turbocharger capable of producing 0.27 MPa of boost, and Engel et al. [3] obtained 320 kW by the 5.670 cm3 GM LS1 * Corresponding author. Tel.: +39 049 8276746; fax: +39 049 8276788. E-mail address: [email protected] (M. Masi). 0306-2619/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2009.09.033
V8 engine adapted for the ASA Stock Car racing series following a strategy of cost savings. Modifications to the intake and exhaust systems aimed at improving volumetric efficiency are very effective to gain specific power. Many standard books in the literature report theory and experimental data about aerodynamics of engine head manifolds, (see, e.g., Heywood [4] and Stone [5]). Since the 1960s the research of Kastner et al. [6] has well established the effects of poppet valve characteristics on engine performance. Woods and Khan [7] and Annand and Roe [8], among others, are classical works dealing with the head manifolds discharge coefficients. A big effort was done to study acoustic waves effects into intake and exhaust systems in order to improve global performance of internal combustion engines. Winterbone and co-workers reviewed a century of researches on the analysis and design of internal combustion engine breathing systems [9]. A review on design and simulation of high-performance engines is due to Blair [10]. Many insight about the methodological approach to design and performance estimate of motorbike high speed racing engines are given by Boretti [11] and related references, and recently by Mattarelli [12]. While theory and practice on engine testing are intrinsically interconnected to the advances in internal combustion engines technology (see, e.g., the classical text by Judge in the 1950s [13] or the more recent by Plint and Martyr [14]), only few references exist about testing of high speed racing engines. Among these, Harrison and Dunkey [15] study 600 kW V10 F1 engines of the early 2000s to develop a decoupled hybrid method
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Nomenclature reference area AR BBDC ABDC before and after bottom dead centre BTDC ATDC before and after top dead centre CA crankshaft angle discharge flow coefficients Cd, Cf EO, EC exhaust valve open, exhaust valve close IO, IC inlet valve open, Inlet valve close, moment of inertia at the crankshaft of all the engine Ieng moving components moment of inertia at the crankshaft of the whole driveIdl line moment of inertia at the crankshaft of applied load IL um mean piston speed um° mean piston speed at maximum power (from steadystate tests on the modified engine clutched on 4th gear) bmep brake mean effective pressure
calculating wave dynamics in the intake system of single cylinder racing engines. They perform steady-state tests on single cylinders similar to those of the real racing engine, as some F1 manufacturers do in order to reduce the costs involved in engine testing. In spite of the advantages of this testing criteria, the authors clearly report the difficulty in reproducing the V10 engine performance by means of single cylinder tests. The reason of this are the unsteady cyclic nature of the intake flow, the friction losses per cylinder, which are higher in the single cylinder engine due to extra mechanical systems required to balance the engine, and the absence of exhaust tuning effects due to the common exhaust of the cylinders in the same bank. To account for a more realistic contribution of the exhaust gases blow-down, Blair suggests to test the complete racing engine by the slow speed ramp testing mode technique [16]. This old technique consists in measuring the main engine performance parameters at fixed throttle position while engine rotational speed is increased within a range in a relatively long time (10–12 s). It was adopted among the others by Weslake in the 1950s to develop race engine designs and is widely used today by technicians involved on high speed engine tuning despite of the lower accuracy associated with the testing mode. In spite of these theoretical and methodological references, the availability of measured data about very high specific power internal combustion engines for racing motorbikes is not so high. In his renowned book [17], Blair provides very useful empirical guidelines for designers and some interesting experimental data about the performance of high speed engines, including those about the V-twin 955 cm3 FIM Superbike Ducati engine from the work by Boretti et al. [18]. This work presents the results of an experimental analysis used to validate numerical calculations aiming at predicting engine performance. The experimental data are shown as nondimensional ratios with respect to maximum torque and power, but engine mean effective pressure is known to be higher than 1.4 MPa. The same scaling technique of the experimental data is used in Cantore and Mattarelli [19] about the performance of a 2003 prototype MotoGP three-cylinder inline engine. A methodology for a parametric design based on similarity rules is developed, and the performance curves of 1.3 MPa mean effective pressure MotoGP engines, obtained by 1D gas-dynamics numerical simulations of different engine layouts, are presented. These curves include a four-cylinder inline configuration, the same of the engine that is studied here. This paper presents an experimental study [20] of a high speed stock motorbike engine [21], which was modified in compliance
bmep°
dLmep p0, T0 pT R T TL Tdl W WOT
brake mean effective pressure at maximum power (from steady-state tests on the modified engine clutched on 4th gear) driveline absorption in terms of mean effective pressure upstream throttle stagnation pressure and temperature throttle pressure gas constant engine torque at crankshaft load torque absorption driveline torque absorption mass flow rate wide open throttle
Greeks
c x
specific heat ratio rotational engine speed
with the 2007 FIM Road Racing Supersport Italian Championship (CIV) [22] to gain specific power. After a synthetic description of the engine modifications, the attention manly focuses on: – The experimental results about the aerodynamic performance of the modified intake and exhaust ducts inside the engine head, involving discharge coefficients measurements on a steady-flow test rig obtained by modifying an industrial fan test rig [23]. – The reliability of the slow speed ramp test performed on the chassis dynamometer as a suitable testing technique for racing engines. – Brake torque, power and specific fuel consumption measurements both at wide open throttle and at part load. – Power losses measurement due to driveline end differences between the engaged gears. Finally the effects of some valve timing parameters on performance are also evaluated. Presented results show torque and power increases for the modified engine of 16% and 33%, respectively. The bike competed in the final part of the 2007 Supersport Italian Championship, scoring the 10th position in the overall standings, the best results obtained being a 4th and a 6th position. 2. The engine The engine powering the Kawasaki ZX-6R/07 motorbike [21] is modified to comply with the standards imposed by the FIM Road Racing Supersport Italian Championship (CIV) [22], that is the racing class for medium volume motorbikes derived from stock production. Table 1 reports the engine technical specifications that are essential for the experimental work developed in the paper. The modifications to the original engine, which are briefly summarised in the following, are all aimed at improving output power within the limitations imposed by the regulations. Combustion chamber shape and volume are modified by altering cylinder head in order to raise the compression ratio (special gasoline is adopted to avoid detonation). Fuel injection system ECU is substituted, and ignition advance and injection timing and duration maps are experimentally tuned as functions of regime and load. Camshafts are changed, in particular cam profiles are shaped to better perform at the higher rotational speeds while valve lift is
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Table 1 Engine technical specifications. Engine model type architecture
Kawasaki ZX6R-07 4 stroke spark ignition inline 4 cylinders
Unmodified stock engine data Swept volume Stroke/bore Combustion chambers Firing order Intake valves Exhaust valves
Valve train Injection system
599 cc 0.634 4-valves pent-roof type 1–2-4–3 from left to right Outer seat diameter 27.5 mm; lift 8.3 mm; lash 0.2 mm Outer seat diameter 23.1 mm; lift 7.5; lash 0.25 mm.
Cooling
Direct drive, flat valve lifters with DOHC chain driven, helicoidal springs Multi point port fuel electronic injection; 2 racks of 8 holes injectors (rack Keihin in the airbox, passive up to 7000 rpm, rack Mitsubishi in the intake ports gradually shut off over 7000 rpm) Liquid cooled forced circulation by radial pump crankshaft geared
Engine parameter/component
Value/intervention
Modifications at the stock engine Compression ratio Fuel Equivalence ratio ECU Piston speed Camshaft Valve springs Valve timing Exhaust system Cooling system Lubricant
14.5 WLADOIL unleaded racing fuel (M.O.N. 89.1 ± 0.4) Ref.: WLADOGAS BB1 Cod. 22544600 1.11 MoTeC m800 22.66 m/s Changed Changed IO: 48° BTDC; IC: 82° ABDC (310° total crank angle opening) EO: 71° BBDC; EC: 39° ATDC (290° total crank angle opening) Substituted with the Akrapovic S-K6EFT5T-ACT kit Supersport 2007 Substitution of the original heat exchanger with a unit able to dissipate a maximum thermal load of about 80 kW SAE 5 W-40 oil
kept constant. Valve springs are changed as well as a necessary consequence. No changes in mechanical components are allowed to reduce mechanical friction, so the only way to improve organic efficiency, apart from surface finishing, is to use a better lubricant. The original coolant heat exchanger is substituted with another unit that is placed in the same location but has a surface that is three times as much. The two following subsections deal with the air breathing system components and the modifications applied to them in order to increase the engine volumetric efficiency. 2.1. Intake and exhaust systems The upper part of Fig. 1 shows the path of the fresh charge inside the air intake system that supplies air to all the cylinders. For sake of simplicity this system is subdivided into three main subsystems: the secondary pipe, the air filter box, and the primary manifolds. The secondary pipe (‘‘airscoop” in Fig. 1) features a large diffuser that is split into two rectangular ducts crossing the steering zone and benefits from ram charging at high speed due to its inlet on bike front. The air filter box (‘‘airbox” in Fig. 1) is a 7.5 l plenum between the secondary pipe and cylinder intake runners. Each of the four independent primary manifolds is made of a short bellmouthed duct (‘‘intake trumpet” in the lower part of Fig. 1), a throttle body equipped with two throttle valves in one throttle bore (a secondary ECU controlled valve is placed upstream the primary one which is used for load control), and the two port passages inside the engine head. Only slight modifications were made to the intake system layout upstream the ports in the engine head, namely:
(1) the bell-mouthed ducts were shortened to rise their tuning frequency, (2) the secondary throttle valves were kept at their maximum opening, as their role is not significant at the rotational speed range of interest, (3) the original air filter was substituted with a 3 mm sponge filter to reduce the pressure loss (although a minimal benefit was achieved). The exhaust system beyond the engine head is completely substituted according to FIM regulations [22], maintaining nothing but the location and number of the original silencers. The exhaust piping downstream engine head is assembled from steel tube components and presents the classical ‘‘4-into-2-into-1” configuration (see Fig. 2). The two intermediate secondary branches collect the primary exhausts of cylinders 1–2 and those of cylinders 3–4, respectively, and discharge the exhaust gases through an Y-junction in the common tailpipe that ends in the silencer. This layout enhances power output at the high speed range of engines with 1-2-4-3 firing order [24]. In addition, Fig. 2 shows two ‘‘interference” pipes (16 mm diameter, and 35 mm and 40 mm length, respectively) connecting the primary manifolds before their Y-junction with the secondary branches. Table 2 reports the main dimensions of the engine air breathing system. The silencer (see the lower part of Fig. 2) is of absorption-type, because of its low exhaust counterpressure [25]. It is made of a 1 mm external steel jacket (section of 13341 mm2) filled with an absorbent lining that covers a 380 mm steel pipe (diameter 54 mm, thickness 1 mm) with a perforated central part (length
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Fig. 1. Intake system (above) and details of the external air section (below).
Fig. 2. Exhaust system: exhaust line (upper-left), details of the interference pipes (upper-right), silencer (lower-left) and its simplified scheme (lower-right).
273 mm). Holes diameter and spacing are about 2 and 1.5 mm, respectively.
isentropic steady-flow discharge coefficient Cd, defined for non chocked flows as:
2.2. The head and the discharge coefficients measurement
" 1=c ( ðc1Þ=c #)1=2 WðRT 0 Þ1=2 pT 2c p Cd ¼ 1 T c1 AR p0 p0 p0
The shape of intake and exhaust ports through cylinder head are modified and piping internal surface has been polished to improve its aerodynamic performance. This can be expressed in terms of the
where W is the actual mass flow rate, p0 and T0 are the upstream pressure and temperature stagnation condition, pT is the static pressure just downstream of flow restriction, g is the specific heat ratio,
ð1Þ
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M. Masi et al. / Applied Energy 87 (2010) 1641–1650 Table 2 Main dimensions of external air section and exhaust system. System
Subsystem
Intake
Secondary pipe (airscoop)
Ram intake Steering zone
Air filter box (airbox)
Zone A
Zone B Zone C
Exhaust
Primary exhaust manifolds
Secondary branches Tailpipe
Length (mm)
Equivalent diameter (mm)
Volume [l]
0 280 0 140 0 95 215 – 0 50
67.8 98.37 69.56 (2) 87.4 (2) 123.62 123.62 190 – 105.25 (2) 108.82 (2)
–
28.7 30 38 38 40 40 43 43 54 54
– – – – – – – – – –
Cyl. 1
Cyl. 2
Cyl. 3
Cyl. 4
0 10 145 445 0 180 0 160 250 1070
0 10 150 475
0 10 155 465 0 180
0 10 140 455
and AR is a reference section that can be defined in several different ways (see Heywood [4]). The geometry of the modified ports (see Fig. 3 for the main dimensions of the equivalent rectified geometry) is acquired by applying a laser scanning technique to a rubber ‘core’ obtained by filling the empty space within the ports. Surface data are then handled with the RapidformÒ 3D Scanning Software (see Fig. 4) to get area measurements at various cross-sections. The aerodynamic performance of the modified intake and exhaust ports is measured on an experimental apparatus derived from a type A industrial fan test rig that is built according to the UNI 10531 Italian Standard [26], which is equivalent to the ISO 5801 Standard. The main scheme of the original test rig, described in details by Martegani et al. [23], is made of an auxiliary fan that blows into a test chamber on the suction side of the fan to be tested. In this type of layouts, an orifice plate is
– 3.6
3.0 0.9
provided by the standard for flow rate measurements. The facility has been modified by substituting the auxiliary fan with a variable speed fan that generates a negative relative constant pressure inside the plenum chamber. A one-cylinder slice of the engine head replaces the fan to be tested. In this way, a constant pressure drop is kept between ambient pressure and plenum chamber pressure across the opened ports, so that the air flow rate can be measured as a function of valve lift only. Dry and wet-bulb ambient temperature and the temperature inside the plenum chamber are measured using mercury thermometers. Barometric pressure is measured using a mercury barometer, whereas a water manometer is used to measure the pressure in the plenum chamber. The orifice plate differential pressure, that is needed for flow rate determination, is measured by a differential water manometer. The accuracy estimated by the Kline–McClintock criterion [27] is less than 1.3% for all the presented discharge coefficient measures. Both direct and reverse flow discharge coefficients are experimentally obtained for different values of the pressure ratio across the ports. Fig. 5 illustrates a sample of the discharge coefficients values of intake and exhaust ports at a fixed pressure ratio of 1.035. In particular, Fig. 5a shows the values of the discharge coefficients Cd that is normalised with respect to a fixed reference section AR (the area of the valve inner seat) as proposed by Woods and Khan [7]. Fig. 5b shows the values of the discharge coefficients Cf that is obtained adopting the ‘‘valve curtain area”, which is a linear function of the valve lift, as reference section AR [8].
duct area / piston area
0.60
3. Engine test rig 0.50 0.40
exhaust intake
C
B
A
0.30 0.20 0.10 0.00 0
unchanged until 1sth Y junction
engine head
40
80
120
160
200
240
distance from valve [mm] Fig. 3. Primary intake (above) and exhaust manifolds (below): sketch and main dimensions.
The engine is tested in an acoustic-insulated cell (Fig. 6) equipped with an intake air fan, which supplies air for engine cooling and breathing (flow rate: 0.37–1.32 m3/s; total pressure: 1176– 1715 Pa), and an exhaust extractor, which removes combustion gases from the room. Inside the cell, a Dynojet Mod. 250i chassis dynamometer (flywheel inertia allows 373 kW maximum power absorption) is coupled with the Jacobs Mod.12JC30 air-cooled eddy current dynamometer and a 1162.5 Nm torque cell. The instrumentation also include sensors for ambient air temperature, pressure and humidity. The results of the tests (which are governed by an elec-
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Fig. 4. Picture and digital acquisition of intake (left) and exhaust (right) pipes inside the engine head.
a
0.8 0.7 0.6
Cd
0.5 0.4 0.3 Cd intake Cd intake reverse Cd exhaust Cd exhaust reverse
0.2 0.1 0.0 0
0.05
0.1
0.15
0.2
0.25
0.3
0.35
0.4
lift / inner seat diameter
b
0.8 Cf intake Cf intake reverse Cf exhaust Cf exhaust reverse
0.75 0.7 0.65
Cf
0.6 0.55 0.5 0.45 Fig. 6. The engine test cell.
0.4 0.35 0.3 0
0.05
0.1
0.15
0.2
0.25
0.3
0.35
lift / outer seat diameter
The test rig, both in transient and in steady-state mode, measures wheel performance. This means that the crankshaft engine torque T can be calculated only if all the terms on the right hand side of the following equation are known:
Fig. 5. Direct and reverse isentropic discharge flow coefficients for the intake and exhaust ports defined by different reference sections: (a) the area of the valve inner seat and (b) the ‘‘valve curtain area”.
T ¼ T dl þ T L þ ðIL þ Idl þ Ieng Þ
tronic module) are instantly corrected following the ISO 3046-1 Standard [28] before real time display and storage on a PC. Both steady-state and transient measurements can be performed in this facility. Engine speed measurement is derived from the ECU pulse sent to cylinder 1 ignition coil and its resolution is 1 rpm. According to the manufacturer, the accuracy on other quantities related to the chassis dynamometer are: timing accuracy of 1 ms; drum speed accuracy of 1/100 mi/h; rpm accuracy 1/10 rpm. Torque accuracy can be estimated as 0.2% of the torque full scale reading, that is ±2 Nm, in the steady-state tests carried out using the electric dynamometer. Fuel consumption is obtained by a weight measure. The fuel tank is arranged on a load cell with a resolution of 1 g and a maximum linearity error of 2 g. The acquisitions are completed by the air–fuel ratio measurement, derived from the signal of the lambda sensor.
where x is the rotational engine speed, TL is the applied load, IL, Idl and Ieng are the moments of inertia of the applied load, driveline and engine, respectively. Thus, in order to obtain the brake performance curves, it is necessary to determine the driveline losses Tdl induced by the gearbox and the transmission (this is necessary for the simpler steady-state tests (dx/dt = 0) as well). To this end, a ‘‘coast down” test is carried out (driveline inertia Idl must be known or estimated), in which the engine of the geared bike is first brought to maximum rotational speed, then the clutch is disconnected manually (T = TL = Ieng = 0) and torque is acquired during the free deceleration of the complete driveline led by the inertia of the chassis dynamometer flywheel. The description of performance curves at a fixed throttle position for a specified engine setup by traditional steady-state tests requires at least 20 acquisitions at different rotational speeds, so the engine works for no less than 30 min roughly. As a consequence, a single racing engine, which is optimised for a few hours
dx ; dt
ð2Þ
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4.5%
4.5% dLmep: experiments dLmep: data fitting dLmep*um: experiments dLmep*um: cubic regr. data fitting
dLmep / bmep°
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[dLmep x um] / [um° x bmep°]
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um / um° Fig. 7. 4th Gear driveline losses.
a
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bmep: sweep test bmep: steady-state test
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operation, cannot bear several full load tests at many rotational speeds. Other researchers (see, e.g., Blair et al. [16]) suggest to overcome this limitation by changing the testing approach for racing engines and carrying out an experimental campaign mainly based on short transient runs, i.e. slow speed ramp tests briefly called ‘‘sweep tests”. The complete performance curves (except
80% 70% 60% engine 1
50%
engine 2
40% 30%
40%
50%
60%
70%
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100% 110%
um / um° Fig. 9. Steady state test mean effective pressure measurements (4th gear) for new engine (engine 1) and worn engine (engine 2).
for the data about specific fuel consumption) are obtained from a 10 s-to-12 s sweep test. Regardless of measurement uncertainty, which is due to the estimation of Ieng and Idl in Eq. (2), this testing mode is certainly the least stressing for the engine. The opportunity to perform the major part of the tests in sweep mode, even if its reliability is lower than the more precise braking mode, is probably the only way to test different setups on such an engine at the same wear level in all tests (see the discussion about Fig. 9). To evaluate the repeatability of the measurements, all the slow speed ramp tests are repeated at least three times. Because of the reduced number of test repetitions, accuracy is determined by means of the Student criterion with a 90% confidence, as established in the European Standards [29]. Repeatability errors in torque measurements for the engine revving range shown in all the presented figures are under 8% in the worst case and the mean error is under 2% in almost all tests (the worst case mean error value is below 4%, this limit being claimed as typical when testing highperformance engines [16]; see the discussion below about Fig. 8a). Most of the tests are performed with the engine clutched on 4th gear, to guarantee operating stability in almost all the rotational speed range that is useful during a competition. 4. Experimental results
um / um° 105%
110%
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100%
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40% 30%
40%
50%
60%
70%
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90%
[bmep x um] / [bmep°x um°]
b
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um / um° Fig. 8. Brake mean effective pressure and power per piston area measurements (4th gear): (a) comparison between sweep test and steady-state test at WOT and (b) comparison between the stock and the modified engine.
This section presents the experimental results about the engine performances in its racing setup. Although torque and rotational speed are measured directly in the test rig, all the experimental results in the present paper are shown in terms of brake mean effective pressure (bmep) and piston speed (um), and are made nondimensional using the factors bmep° (1.38 MPa) and um° (20.6 m/ s), which are the brake mean effective pressure and the piston speed, respectively, measured by a steady-state mode test in which the engine operates in its racing setup at maximum power clutched on 4th gear. Fig. 7 shows driveline losses with the 4th gear engaged in terms of mean effective pressure (dLmep) and of power per piston area (dLmep um). Both non-dimensional curves are obtained from a cubic regression fitting of mean effective pressure measured data. As stated in the previous section, some test are carried out in steady-state mode to compare the results with sweep-mode measurements and to estimate the accuracy level of the latter. The curves in Fig. 8a show the wide open throttle (WOT) performance for the engine in its racing setup as obtained by the sweep-mode tests. Brake mean effective pressure and power per piston area values at WOT condition as measured in steady-state mode tests are
M. Masi et al. / Applied Energy 87 (2010) 1641–1650
plotted as dots in the diagram as well. The error bars of the sweep test curve show the estimate of the random instrumentation errors only (see Plint and Martyr [14]). It is clear that the agreement between the results obtained with the two test modes is not complete, especially at low rotational speeds. While differences at low regimes are strongly affected by engine instability, sweep measurements also suffer by the rough evaluation of the overall rotating part inertia Ieng and Idl in Eq. (2) (it was not possible to perform massive tests because of motorbike availability and the need of preserving against wear an engine ready for competitions). Moreover, exhaust pipe wall and exhaust gas temperatures, and then wave propagation velocities, are not identical in the two test modes [30]. However, although the error band is quite large, the diagram shows that the results of the sweep test mode are well inside the band of accuracy of the traditional steady-state test for almost all the rotational speed range of practical interest. To enforce the advantages deriving from the sweep test technique, the performance of two identical engines were compared in Fig. 9. The former, numbered as engine 1, is the one used in this work, the latter, numbered as engine 2, worked in competitions for about 1000 km, undergoing a number of working cycles comparable with those of an extensive test bed experimentation based on steady-state tests. Despite of the amplitude of the uncertainty, the tests on these engines performed in the steady-state mode clearly quantify the mean incidence of wearing on torque as 2% of the maximum brake mean effective pressure, with peaks that are greater than 5% at some revving speeds. Fig. 9 also shows clearly that the reliability and the duration of the modified engine are much worse. This drawback is however emphasized by the technical regulations [22] that do not allow the substitution of many mechanical components which would require different designs and materials to keep up with the increased performance. In Fig. 8b the brake mean effective pressure and power per piston area curves of the modified engine described in Section 2 are compared with the corresponding performances data measured for the stock engine. The sweep-mode tests highlight a consistent performance gain for the modified engine at all the regimes above 45% of um°. Moreover, the increment in maximum brake mean effective pressure for the modified engine is about 16% (from 1.22 to 1.42 MPa), while maximum power per piston area shows an improvement of more than 33%. The comparison between the two engine configurations shows that the bmep curve maintains approximately the same shape, but, as expected, the trends are sharpened with higher peaks and lower drops for the modified configuration. The most substantial modifications to the air intake system and the total replacement of the exhaust system result in an improvement of the volumetric efficiency which occurs at the higher rotational speeds in particular. Performance is therefore enhanced at high regimes, whereas it drops significantly at lower regimes. This may be viewed as a potential drawback for stock engines, but racing engines rarely operate below 60% of their maximum power rotational speed. Fig. 10 shows the specific fuel consumption evaluated in a steady-state test in which the time required to use up to 100 g of gasoline is recorded, the uncertainty of this indirect measure being estimated using the Kline–McClintock criterion [27] as suggested by the standard [29]. The curve shows some peaks (the highest occurs at about 50% of um°) at the regimes featuring volumetric efficiency drops due to the increasing in pumping work. Note, however, that the magnitude of these peaks is not very significant if the scale on the ordinate axis is considered. A full comparison with similar high technology racing engines is difficult, due to the little availability of experimental data. A paper by Cantore and Mattarelli [19] reports some non-dimensional experimental data about torque and specific fuel consumption for a three-cylinder MotoGP year 2003 prototype. While the
340 Specific Fuel Consumption
330 320 310
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um / um° Fig. 10. Specific fuel consumption of the engine.
three-cylinder MotoGP engine offers appreciable differences in specific power, and more important, in power per piston area (they are about 8% and 14% greater than those of the engine tested here), the performance curves exhibit some agreement about the location and the relative amplitudes of peaks. This is because the air breathing system layout, except for the telescopic taper device that can vary continuously intake pipe length, is mostly independent of the technological level within the very high speed engines category. Note that the brake mean effective pressure of the engine tested here is about the same of those tested in [18] and [19]. Thus the results presented for this SuperSport racing motorbike engine have a general validity for very high speed motorbike engines including the MotoGP ones, despite to their higher technology level (comparable to that of F1 engines [19,12]). Fig. 11 illustrates the WOT brake mean effective pressure and the increment in transmission losses (in terms of mean effective pressure) when the 6th gear of the driveline is engaged instead of the 4th. The average of the difference in mean effective pressure absorption at all the rotational speed is 1.8% of bmep°. Fig. 12 illustrates brake mean effective pressure performance at different loads for the driveline with the 6th gear engaged. The dotted line in the diagram is the 100% full throttle (WOT) brake mean effective pressure curve obtained from other tests carried out with the driveline 4th gear engaged. The addition of the average of the difference in mean effective pressure absorption between 4th and 6th gear (Fig. 11) confirms the previous rough estimate of this value and indirectly confirms once again the acceptable degree of repeatability for the measures. As expected, part load curves show the performance falling down because of the deterioration of the volumetric efficiency, which is strongly affected by the throttling imposed by load control. Please note the trend at the medium to low rotational speed where the 50% full load torque curve exhibits performance data that are better than those of the 75% torque curve. This behaviour is often featured by engines the valve timing of which is optimised for high rotational speeds (e.g., racing engines), where a strong fresh charge backflow in the intake manifolds occurs at the beginning of the compression stroke for low regimes. In such engines the throttling by the load control system at low regimes acts like a barrier to backflows, and it is not unusual that further throttling increases brake mean effective pressure. This is the main reason of the secondary throttle valve in almost all the stock engines, as confirmed by Wakimura and Nitta [31]. On the other hand, the secondary throttle has been locked wide open in these tests because of the high operation regimes of the engine. Finally, Fig. 13 shows non-dimensional brake mean effective pressure and power per piston area curves obtained with an ad-
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110%
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um / um° Fig. 11. Brake mean effective pressure on 6th and 4th gear and difference in mep absorption between gears (WOT).
110% 100%
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5. Conclusions An experimental investigation on the performance of a motorbike high speed racing engine was presented. The modifications to the stock engine and the geometry of the complete air breathing system, in compliance with the FIM Supersport race class regulations, were illustrated and experimentally tested. Results show a gain of about 16% in maximum torque and 33% in maximum power for the modified engine. A comparison with the performance of an identical engine having expired its short racing life cycle pointed out that massive steady-state runs at the test bed could not be performed because of performance degradation due to wearing, which is of the same order of magnitude of the performance improvement generated by specific design changes. Thus, engine performance at both wide open throttle and part loads, intake valves timing effect, driveline and gearing influence were analysed by slow speed ramp testing mode on a chassis dynamometer test bench. The worst case mean error value on torque measurements is below 4%. The comparison with steady-state measurements and the error analysis revealed an acceptable degree of accuracy for the slow speed ramp testing mode even for a low number of test repetitions, and, therefore, suggests this testing mode is suited to high speed racing engine in order to save wearing, costs and time. Finally, the indications provided here can be shared among different designs by using similarity rules, and the presented measures both for discharge coefficients and gross performance offer useful data for one dimensional gas dynamics engine simulations.
The authors gratefully acknowledge Giorgio Mozzo for the great tuning engines experience he shared in this work, and Andrea Mozzo for the great help in the dynamometer tests.
bmep*um MOP 107°
40% 40%
imum gain in brake mean effective pressure (and power per piston area) of about 4% in the range from 48% to 55% of um°, with a little decrease at the maximum power per piston area regime (however, the difference of about 1% is comparable with the local measures uncertainty). This result suggests that additional modifications to the camshafts may be beneficial.
Acknowledgments
[bmep x um] / [bmep° x um°]
bmep / bmep°
Fig. 12. Brake mean effective pressure at different loads (6th gear).
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um / um° Fig. 13. Brake mean effective pressure and power per piston area curves for MOP at 104° ATDC and 107° ATDC.
vance of intake valve mean opening period (MOP) angle from 107° after top dead centre (ATDC) to 104° ATDC. The advance, which anticipates intake valves closure, is made by an azimuth shift of the intake camshaft. Brake mean effective pressure is improved in almost all the medium/high range of regimes, obtaining a max-
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