ANNALS of Faculty Engineering Hunedoara International

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ISSN 1584 2665 print ISSN 1584 2673 online, exhaust collector pipe the second with a 785 mm pipe The inner diameter was 56 mm that connected to the original collector that. had 50 mm inner diameter This diameter difference proved to be useful as it generated small but significant pressure waves that. lead to the final design, 4 RECORDING THE DATA NECESSARY TO VALIDATE THE MODEL. For registrating the engine parameters that can be used to validate the model a rolling road dynamometer was used The actual. measurements were conducted on a Superflow CycleDyn Pro SF 250 WynDyn 3 2 dynamometer It has a special feature namely. that the tested vehicle s air intake openings are supplied by air at a speed that would exactly correspond to the road speed of the. vehicle under the test This reproduces the actual road conditions and increases the accuracy of the test During the tests it was. considered to be fundamental to record parameters that could be compared to parameters produced by the simulation program. These were power torque exhaust gas temperature in the header pipes air fuel ratio lambda wheel slip parasitic losses in the. driveline of the vehicle Table 2 comparison of published and measured engine parameters. To get accurate information on the state of the engine to be Parameter Published manufacturer data Measured data. tested the first recordings were taken with the OEM exhaust Power 54 7 kW 8800 rpm 54 05 kW 9250 rpm. Torque 64 Nm 7000 rpm 58 5 Nm 8750 rpm, system in place The collected data was corresponding well to. the parameters published in the vehicle s service manual 7 The comparison of these parameters can be found in Table 2. The possible reason for the differences is the usually generous description of the vehicles parameters by the manufacturers which. is a well known marketing strategy Aside from that the test engine was assessed to be in reasonable condition to continue the. validation work 60 0 120, After this stage the OEM exhaust silencer was removed P 550 mm. P 785 mm 110, and replaced with the straight through pipes As can be 50 0.
seen in Figure1 none of the replacement pipes improve on 100. the peak power figures but considerably affected the M 785 mm 90. midrange torque unfortunately not for the better This M Std Kip. decrease in torque is worst in the range between 4200. 6000 rpm and is highlighted most with the 785 mm 20 0. exhaust pipe,5 STRUCTURE OF THE SIMULATION MODEL 50. The simulation model was built with components 0 0 40. described in the INTRODUCTION section of this present 2000 4000 6000 8000 10000. paper It looks like a network of different elements Figure Engine speed rpm. 2 Data describing the individual parts can be input. through dedicated pop up panels This does not mean any Figure 1 Power and torque characteristics of the engine with the OEM and. problem with pieces that can be measured the two manufactured exhaust systems. easily in reality These are intake and exhaust,pipes connecting rod centre to centre. distance compression ratio etc,To define the Coefficient of Discharge of the. intake and exhaust valves needed special,equipment The highest pressure ratio of this. apparatus was 1 1 but the actual pressure,ratios occurring across valves of internal.
combustion engines are much higher than,this causing flows reaching sonic conditions. locally Since the value of Cd is affected by the,velocity of flow the recorded discharge. coefficient values were extrapolated up to a,pressure ratio of 2 using a built in target. Figure 2 Outlay of engine simulation model,function of the simulation software 6 The. returned result was a surface of Cd values as a function of relative valve lift and pressure ratio Figure 3. All valve systems have two Cd surfaces one for normal direction of flow and one for reverse flow Normal means inflow for intake. valves and outflow for exhaust valves while reverse means outflow for intake valves and inflow for exhaust valves considering the. 84 Fascicule 2, ANNALS of Faculty Engineering Hunedoara International Journal of Engineering.
cylinder as datum These later Cd maps in the reverse direction define the amount of fresh charge escaping from the cylinders. during blow back through the intake valve while at the exhaust valves the reverse Cd maps describe the amount of exhaust gas. recirculated internally in the engine, Other important parameter of valve systems is the opening and closing. point of the valves relative to the momentary piston position and the. actual time areas of the opening processes Input of these parameters was. only possible after the accurate measurement of the camshafts The details. of the measurement can be found in the literature 3. The cylinders are the central parts of the model Beside of the obvious. measures and relative ignition sequence the frictional mean effective. pressure FMEP needed to be given too The internal friction Figure 3 Extrapolated map of Discharge Coefficient. characteristics of the engine could not be recorded directly only the values as a function of relative valve lift and pressure ratio. obtained during the dyno test were available These were masked heavily across the valve. by the driveline losses of the vehicle therefore these were used only as base for the actual data input during the simulation runs. There is a debate in the literature on the exact shape of the frictional losses diagram Blair 1 suggests linear increase of losses with. increasing engine speed while Yagi et al 8 10 conducted researches based on motorcycle engines and suggests progressive. increase with speed Since the test engine is a,motorcycle engine the actual values used in the. simulation were chosen based on the data presented in. the works of Yagi et al resulting in progressively. increasing FMEP values of 0 41 bar 1000 rpm and 1 8. bar 10000 rpm P 785 mm Modell,6 RESULTS OF MODELLING. After setting up the model the fourth run produced 2000 4000 6000 8000 10000. acceptable results that quite well matched the Engine speed rpm. measured power and torque characteristics of the Figure 4 Comparison of measured and simulated power curves obtained. engine Seven more runs needed to refine the model using the 785 mm exhaust pipe. with the 785 mm collector pipe to reach a maximum, variation of 2 9 kW between measured and simulated 60. power values This equalled an average error of 5, which value is considered to be sufficient precision 2 50.
Figure 4 The torque characteristics followed a similar. nature and the average error stayed within 5 in this M 785 mm Modell. case as well Figure 5 30, As a check the same way as was done in reality the 2000 4000 6000 8000 10000. exhaust collector length was decreased to 550 mm and Engine speed rpm. the model was run again In this case the error value. Figure 5 Comparison of measured and simulated torque curves obtained. between simulation and measurement was 24 at,using the 785 mm exhaust pipe. 3000 rpm but at the most critical 5000 rpm the error. decreased to 1 and never exceeded 7 in average which was. true for the power and also for the torque values, In the torque curves it can be clearly seen that general trends. troughs etc occur at the same engine speed in the model as in. the measurements Based on this general behaviour of the. simulated engine it can be concluded that the model reflects. the real engine parameters with appropriate precision. 7 EXAMINATION OF THE MODEL, During further simulation the object was to find the reason for. Figure 6 Delivery ratio curves for both cylinders, the great decrease in the torque between 5 6000 rpm Using.
the possibilities of the simulation software the Delivery Ratio DR curves were investigated cylinder by cylinder basis This showed. that the rearward facing cylinder which is number 1 considering the ignition sequence has its DR far from ideal Figure 6. 85 Fascicule 2,ISSN 1584 2665 print ISSN 1584 2673 online. To find the cause of the deficit in Delivery Ratio the. animation utility of the simulation software was,used Figure 7 This function made possible to. visually check the pressure waves in motion within. the gases of the engine while also showing the, valves at their appropriate position relative to the. With this method it was discovered that the, shortage of DR of cylinder 1 is caused by a pressure. wave emanating from itself This wave propagates,not only towards the open end of the system but.
enter s all other openings within the exhaust, system In this case it intrudes into the pipe system. Figure 7 Animation showing cylinder 1 at when the exhaust pressure wave. of cylinder 2 as well through the cross connecting. enters the intake port causing back flow and diluting fresh charge with. balance pipe Since this is a V engine the exhaust exhaust gas The red circle shows wavelets produced by a welding protruding. valves of cylinder 2 are closed when the pressure into the exhaust flow Purple shows the superponated pressure blue the. wave arrives there and the wave is reflected back returning leftward waves red is the escaping rightward pressure waves. with the same attitude eg as a compression wave, the same way as it arrived from So it travels again towards the open end of the system but also enters any pipes now to the. header of cylinder 1 It travels up to the valve but arrives there during valve overlap. The result of this is that the arriving compression wave partially blocks the exhaust process and the outflow from cylinder 1. happens at a much lower speed than from cylinder 2 To worsen things the arriving wave pushes the fresh charge back to the. intake port diluting it with exhaust gas Therefore the combustion process is lower efficiency in this cylinder. 8 ACHIEVED IMPROVEMENTS, To solve the problem caused by the badly timed reflected wave blocking the balance pipe totally seemed to be a good idea It really. would flatten the torque curve at 5000 rpm for a price of loosing around 5 kW from the peak power As an alternative solution the. length of the intake tracts were changed but this would only replace the depression with the same amount of lost kW s from peak. power value, To get the most horizontal torque characteristics a collector pipe shape was found in the simulation environment that contains a 80. mm long conical section where in reality the 6 mm step in diameters was between the OEM collector and the manufactured end. pipe Continued search for the flattest torque curve lead to a solution where the pressure waves originated from cylinder 2 are. used to help the gas exchange process in cylinder 1 This was achieved by an 860 mm collector pipe As a result the dip in the. torque characteristics could be decreased and the cylinders now can work with the least differing amount of air. To arrive to this design about 130 simulation runs were conducted During this work it became clear that the engine could not. respond to any changes to produce not only better torque and power distribution with fixed peak values but to improve on the. peak figures too It was observed that modifications aimed at the physical dimensions of its gas exchange system were not enough. to produce the required result After almost all solutions tried and excluded the reasons were pinpointed to be the inadequate. time area values for the valve events At this point three possibilities were evaluated. changing the intake valve time area,changing the exhaust vale time area.
changing the time area of both valves, On the grounds of getting the most out of the engine camshafts should have been changed This would have been essential in. achieving the desired valve lift and duration characteristics As described in 3 the intake valve springs are at their limit at full valve. lift therefore increasing the intake valve lift any further would require very expensive modifications For that reason changing the. timing and lift values on the intake side or both sides together were not tested. In contrary to the intake side the exhaust valve spring has some margin at full lift for some improvements Therefore exhaust valve. open duration was lengthened by 22 Crank Degrees CA and the lift were increased by 1 3 mm This modification immediately. returned an overall improvement of power and torque in every aspect. Checking on the p V diagram Figure 8 a b the reason of the improvement could be easily identified The area in the diagram. corresponding to the work consumed by the exhaust stroke has been decreased in both cylinders This change also improved the. engine volumetric efficiency and proved to be one of those rare occasions when improvement in performance figures does directly. 86 Fascicule 2, ANNALS of Faculty Engineering Hunedoara International Journal of Engineering. not translate to increased fuel consumption In fact rather the opposite is true To get the same road performance less fuel is. needed thanks to the improved efficiency of the engine In cylinder 1 the effects of pressure wave tuning achieved with the 860. mm tapered exhaust pipe could also been identified Because of the positive simulation results a decision was made to implement. these changes in the material world, Figure 8 a b P V diagrams of the gas exchange process showing the decrease in pumping losses in red The green area represents a small. improvement in the strength of the intake stroke realized by the smaller amount of exhaust gas back flow. a Cylinder 1 b Cylinder 2, Given that the engine should be used within a vehicle an absorption silencer Table 3 Main dimensions of the absorption silencer. ANNALS of Faculty Engineering Hunedoara International Journal of Engineering 83 Fascicule 2 Tome XIII 2015 Fascicule 2 May ISSN 1584 2665 print ISSN 1584 2673 online a free access multidisciplinary publication of the Faculty of Engineering Hunedoara 1 L szl KOV CS 2 Szil rd SZAB IMPROVING THE POWER CHARACTERISTICS OF AN INTERNAL COMBUSTION ENGINE WITH THE HELP OF

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