International Journal of Automotive Technology, Vol 11, No 4, pp 447−453 (2010) DOI 10.1007/s12239−010−0055−8 Copyright © 2010 KSAE 1229−9138/2010/053−01 DEVELOPMENT OF ELECTROSTATIC DIESEL PARTICULATE MATTER FILTRATION SYSTEMS COMBINED WITH A METALLIC FLOWTHROUGH FILTER AND ELECTROSTATIC METHODS H J KIM, B HAN, W S HONG, W H SHIN, G B CHO, Y K LEE and Y J KIM * Environmental Systems Research Division, Korea Institute of Machinery and Materials (KIMM), 171 Jang-dong, Yuseong-gu, Daejeon 305-343, Korea (Received May 2009; Revised 18 September 2009) ABSTRACT−A 3000 cc diesel engine attached to an engine dynamo was used to test three newly developed electrostatic Diesel Particulate matter filtration Systems (DPS 1, 2, and 3) under four steady-state engine operating conditions: idle, 2000 rpm with no load, and 2000 rpm under 25% and 50% loads Of the two developed alternatives, DPS and DPS 2, DPS comprises an ionization section, electrostatic field additional section and Flow-Through Filter (FTF), which achieved almost 90% removal of particulate matter (PM) under the engine’s operating conditions, and the efficiency of the FTF was maintained between 20% and 50% Comparing the long-term performance of DPS and DPS (effectively a serial combination of two DPS 2s) with a commercially-available Diesel Particulate Filter (DPF), the DPS and DPS achieved almost the same efficiency for removing PM as the DPF but had significantly improved (75%~90% lower) differential pressure drops KEY WORDS : Diesel particulate matters, Filtration system, Electrostatic precipitation, Flow-through filter, Removal efficiency, Differential pressure drop INTRODUCTION of diesel particles on human beings is known to be related to particle size, and the presence of smaller particles may be linked with more serious respiratory or cardiovascular diseases (HEI, 2002) International environmental institutes have therefore insisted that the regulation of PM in diesel exhaust should be based on the number of particles rather than their mass concentration (Kittelson et al., 1999) To reduce both the mass and number concentrations of PM in the exhaust gas of diesel vehicles, post-treatment systems using ceramic DPF have been commercialized to retrofit selected vehicles and are also on the cusp of commercialization in light and heavy duty diesel automobiles (Park et al., 2006) Wall-flow ceramic filters, which are coated using metal catalysts, have been shown to have a filtration efficiency greater than 90% for diesel particulates, but diesel ceramic filters are still marred by operational problems, such as insufficient reliability and an excessively high pressure drop caused by solid particles that become clogged under low exhaust temperature conditions Furthermore, the excessive heat released and the high thermal gradient that occurs in the filter during their regeneration can lead to mechanical cracking and failure (Park et al., 2006; Cho et al., 2007) The development of alternative ceramic DPF designs has generated considerable interest in the research community In particular, PM control devices manufactured using metallic materials have shown considerable promise in PM emission control technology because these devices Diesel engines enjoy widespread use in heavy duty vehicles due to their superior fuel economy and durability compared with gasoline engines (Monaghan, 2000) Furthermore, with increasing oil prices, the development of post-treatment technologies, such as DPF (Diesel Particulate Filter) and DeNOx catalysts, and the increasing world-wide demand for tighter controls on CO2 emissions to address global warming, diesel-driven cars have become a viable alternative to gasoline-powered personal automobiles in many European countries Nevertheless, despite their lower fuel consumption, longer durability, and lower CO2 emissions compared with gasoline-driven cars, it is well known that diesel engines emit significant amounts of Particulate Matter (PM) and Nitrogen Oxides (NOX) and thus contribute to the overall PM and NOx pollution of the outdoor environment (Yoon and Cho, 2009; An et al., 2006; Jacob et al., 2006, Jacobs et al., 2006; Jeong et al., 2008) The attendant regulation of PM emissions from diesel vehicles is becoming increasingly stringent as the European Union adopts new legislation in this area (e.g., Euro VI and V) Furthermore, all diesel engines, even newly-developed ones, produce similar amounts of ultrafine particles, and it is the presence of these particles that largely determines the concentration of PM in diesel exhaust The harmful effect *Corresponding author e-mail: yjkim@kimm.re.kr 447 448 H J KIM et al exhibit a relatively low pressure drop and avoid the numerous problems and complex structure of DPFs As a result of these benefits, many researchers have addressed the development of PM removal systems using metallic filters Yoon and Cho (2009) and An (2006) studied the design of a metallic foam filter and analyzed its aerodynamic performance, differential pressure drop, and filtration efficiency Their experimental results showed that the filter removed more than 50% of the diesel PM based on mass Other researchers applied a filtration system with metallic Flow-Through Filters (FTFs) to heavy duty diesel vehicles and investigated their diesel PM and gaseous compound removal efficiency The experimental results from these studies showed that mass concentrations of diesel PM were decreased by 50~70% using the FTF, but number concentrations were decreased by less than 50% (Jeong , 2008; Park , 2006; Bruck , 2001; Jacob , 2006) These metallic filters have open flowthrough passages that permit exhaust gases to pass when their PM capacity is exceeded Thus, the pressure drop does not increase dramatically, but their PM removal efficiency based on number concentration is low compared with that of DPFs (Majewski, 2008) To compensate for the weak ability of metallic filters to remove PM, Park (2007) applied a corona charger upstream of a metallic foam filter and thus increased the PM removal efficiency of the system by 10~20% To date, however, few studies have applied electrostatic charging and precipitation to metallic FTF, even though these devices have become fully commercialized in Korea and several European countries In this study, newly designed electrostatic filtration systems consisting of electrostatic devices and commercial FTFs were developed to achieve PM removal performances as high as those of commercial DPFs by using particle charging and an additional electrostatic force on the FTF to enhance the removal efficiency of the standalone filter The pressure drop and PM removal performance of the combined systems were investigated and compared with those of a commercially available ceramic DPF under a variety of operating conditions using a diesel engine attached to a dynamo et al et al et al et al et al et al EXPERIMENTAL APPARATUS AND PROCEDURE 2.1 Electrostatic Diesel Particulate filtration System (DPS) Figure shows the schematic representations of the electrostatic Diesel Particulate filtration Systems (DPSs) used in the study Figure 1(a) shows the two parts of DPS 1: the electrode, which generates unipolar ions and applies an electrostatic force to the front of the flow-through filter simultaneously, and the filter itself The edge electrode (an astral shape with eight legs) of the first section was aligned parallel to the front section of the FTF The length of the section that generated the ions Figure Schematic representation of the electrostatic DPSs showing 1) the sections that generate unipolar ions and apply an electrostatic force on the flow-through filter and 2) a flow-through filter coated with catalysts Figure Experimental setup of the performance tests using the DPSs and the commercially available DPF was 60 mm; the length and diameter of the FTF were 105 mm and 130 mm, respectively Figure 1(b) shows DPS 2, which is also composed of a section that generates unipolar ions and imposes an additional electrostatic force and the FTF In contrast to DPS 1, the edge electrode (the rod with edges), to which the perforated plate was attached to DEVELOPMENT OF ELECTROSTATIC DIESEL PARTICULATE MATTER FILTRATION SYSTEMS 449 impose the electrostatic force on the filter, was positioned perpendicular to the face of the FTF The total length of the electrode was 205 mm, and the length of the FTF was the same as that of the filter in DPS DPS 3, which was designed to increase PM removal efficiency, was a twostage set up consisting of the same unit as the DPS but manufactured with two filters, each of which was half the size of the FTF used in DPS and DPS The total length of this system was 515 mm 2.2 Experimental Methods The DPS devices were connected to the exhaust of a 3000 cc diesel engine with an engine dynamo The characteristics of the devices were investigated by varying the engine operating conditions and comparing their performance with the DPF under the same experimental conditions The test engine that we used was the 3000 cc diesel engine (Model Frontier, Hyundai Motors, Korea) with a maximum torque and speed of 17 kg·m and 4000 rpm, respectively The engine speed was set at idle and 2000 rpm with no load and 2000 rpm with loads of 20% and 50% The experimental set up is shown in Figures and Figure 3, and specifications of the test diesel engine are summarized in Table The high power supply (Max −30 kV/ 10 mA) was connected to devices installed at the center of the exhaust line, and the sampling probes were inserted just before and after the DPSs to measure their PM removal efficiency To minimize variations in the concentrations of the diesel PM caused by gaseous compounds, all of the sampling lines connected to the rotating dilutor (Model MD-19, Matt Engineering, Switzerland) were electrically heated to 200 o C, and the sampled gases were mixed with air at a dilution Table Specifications of the test diesel engine Engine Model Displacement Max Max type torque speed Diesel 17 kg·m 4000 rpm engine Frontier 2957 cc Figure Photograph showing the experimental set up Figure Size distributions of the diesel PM in the engine exhaust for various speed/load combinations ratio of 1:200 A DMA (Differential Mobility Analyzer, Model 3080, TSI, USA) and a CPC (Condensation Particle Counter, Model 3076, TSI, USA) were used to measure the number concentration and size distribution of the diesel particles before and after the test filtration systems were operated The PM removal efficiency was calculated using the following formula, C (1) η =⎛⎝1 – ⎞⎠ × 100 , C Where, η is the removal efficiency (%); is the outlet concentration of particles per cc; and is the inlet concentration of particles per cc In addition, a device that measures pressure (Testo 350M/XL*testo 454, Testo, Germany) was connected to tabs located at the inlet and outlet of the DPSs, and a thermocouple linked to a temperature monitoring system (Model V18, SDD, Korea) was also inserted into the line upstream of the filtration systems to measure the pressure drop and the inlet temperature out in Cout Cin RESULTS AND DISCUSSION 3.1 Performance Test Results of the Two Types of Electrostatic DPSs Figure shows the distributions of the number concentration as a function of particle diameter for the different engine operating conditions The number concentration of the diesel particles increased, and the size distribution shifted to the right, as the engine speed and load increased Most of the particles from the engine exhaust were found in the nuclei and accumulation modes, and were less than 300 nm in diameter The mean diameter of the particles was in the range of 30~50 nm Figure shows the curves of the corona voltage against current at different engine speeds and loads, which were compared for each DPS to investigate the electrical characteristics of the respective systems The curves for both DPSs moved to the left, and the corona current was higher for the same applied voltage when the 450 H J KIM et al Figure Corona current plotted against applied voltage for various speed/load combinations engine speed and load increased In particular, ‘sparkover’ in the DPSs, which indicates the start of the unstable corona discharge, was observed at lower applied voltages when the speed and load increased This result mainly occurred due to the increased temperature of the exhaust gas For the negative corona discharge, as the temperature of the gas increased, the mean free path between the gas molecules increased, and the frequency of the collisions between electrons and the gas molecules decreased significantly Consequently, those electrons that not collide with the gas molecules are electrically forced to the grounded side Thus, high temperature leads to unstable corona discharge (Kim , 2001) In the case of DPS 2, the initial voltage was higher than that of DPS (Figure 5) Because the distance between the sharp edges of the electrode and the grounded side in DPS (see Figure (b)) was wider than the distance in DPS (Figure (a)), the corona in DPS was initiated at higher voltages than in DPS (Hinds, 1999) Figures and show the PM removal efficiency of DPS against the mean PM diameter for various applied voltage/current combinations at different engine speeds and loads The efficiency of DPS with the FTF was 30~80%, which is higher than that of the standalone filter, whose efficiency was 25~50% at idle, 2000 rpm without a load, and at 2000 rpm with loads of 25% and 50% Because both the charging efficiency of the diesel particles and the Figure PM removal efficiency plotted against PM diameter for various applied voltage/current combinations at different engine speeds for DPS Figure PM removal efficiency plotted against particle diameter for various applied voltage/current combinations at different engine loads for DPS et al DEVELOPMENT OF ELECTROSTATIC DIESEL PARTICULATE MATTER FILTRATION SYSTEMS 451 Figure Removal efficiency at various different engine speeds as a function of changing particle diameter for various applied voltage/current combinations for DPS Figure Removal efficiency for various different engine loads at 2000 rpm as a function of particle diameter for various applied voltage/current combinations for DPS electrostatic force on FTF increased as the voltage applied to DPS increased, the efficiency with which the diesel particles were removed was higher than that of the FTF itself Furthermore, at low speed/load combinations, when the temperature of the exhaust gas was low, the particle removal efficiency was higher than that observed under the higher speed/load combinations because of the higher applied voltages and longer residence times at the low combinations Figures and show the PM removal efficiency of DPS against the mean PM diameter for various applied voltage/current combinations for each of the engine speeds and loads In contrast with DPS 1, the efficiency with which the diesel particles were removed in DPS was 30~50% higher than that of the standalone filter, and the removal efficiency was over 85% at temperatures over 250oC and more than 95% at temperatures less than 200oC, which was similar to the efficiency of the commercial DPF As shown in Figure 1, because the direction of the corona discharge of the ionizer in DPS was parallel to the direction of the exhaust flow, the residence times of the particles and unipolar ions in the charging region were shorter, while the direction of the corona discharge in DPS was perpendicular to the direction of flow, and the charging region was wider than that of DPS Thus, the charging rate of the particles in DPS was expected to be Figure 10 Comparison of the PM removal efficiency at the mode diameter (40 nm) among the FTF, the DPS and the DPF at 2000 rpm with a 25% load applied for hours higher than that in DPS 1, which explains why the efficiency of the DPS exceeded that of DPS 3.2 Comparison between the Electrostatic DPSs and the Commercially Available DPF The performance of DPS 2, which showed greater PM removal efficiency than DPS 1, was compared with that of the commercially-available DPF under the same experimental conditions: 2000 rpm under 25% and 50% loads To 452 H J KIM et al Figure 11 Comparison of the differential pressure among the FTF, DPS and the DPF at 2000 rpm with a 25 % load applied for hours Figure 13 Comparison of the differential pressure between the DPS and the DPF at 2000 rpm with a 50 % load applied for hours over 260oC, and the applied voltages and currents in the first and second stages were 13.4 kV/1 mA and 11.5 kV/2 mA, respectively As shown in Figure 12, the particle removal efficiency was over 90% for the whole hours, and the pressure drop (343 mmH2O, Figure 13) was significantly lower than that of the DPF Furthermore, it did not increase over time, unlike the DPF whose initial pressure drop (1700 mmH2O) increased dramatically to 3000 mmH2O after hours of operation CONCLUSIONS Figure 12 Comparison of the PM removal efficiency at the mode diameter (40 nm) between the DPS and the DPF at 2000 rpm with a 50% load applied for hours improve the removal efficiency of DPS at 2000 rpm with 50% load, DPS was designed as a serial combination of DPS 2s, as described in Section 2.2; its performance was also compared with that of the commercial DPF The ceramic DPF used in this study had a diameter of 142 mm and a length of 154 mm The PM removal efficiency was calculated for the peak particle diameter of 40 nm Figures 10 and 11 show the performance of the three filtration systems, the FTF, DPS 2, and the DPF, in terms of their PM removal and observed pressure drop, respectively The temperature of the exhaust gas at 2000 rpm and 25% load was 180oC The efficiency of DPS when the high voltage/current (14.3 kV/1 mA) was applied was over 90%, which was similar to that of the DPF and significantly higher than that of the standalone FTF Furthermore, the observed pressure drop of DPS was 200 mmH2O, which was significantly lower than the 800 mmH2O of the DPF and did not increase during the three hours of testing Figures 12 and 13 show the variation in the performance of DPS and the DPF over hours of the test at 2000 rpm under 50% load The temperature of the exhaust gas was We developed electrostatic DPSs that achieve performance similar to that of commercially available DPFs with respect to PM removal and the pressure drop The DPSs were designed using a commercial FTF combined with an additional electrostatic particle removal technique We conducted performance tests on each system and compared them with those of a commercially available DPF Our major findings may be summarized as follows: The DPSs were tested at constant engine speeds of idle and 2000 rpm, and loads of 25 and 50% at 2000 rpm in a 3000-cc diesel engine The PM removal efficiency of the FTF improved from 20 to 60% to 40 to 95% under the various engine operation conditions using an electrostatic precipitation method The DPS achieved a PM removal efficiency of over 90% at an exhaust temperature of less than 170oC and an efficiency greater than 80%, even at exhaust temperatures above 260oC The one-stage (DPS 2) and two-stage (DPS 3) filtration systems showed similar PM removal to the commerciallyavailable DPF over 3- and 8-hour engine operation at 2000 rpm at loads of 25% and 50%, while their pressure drops were only 200 and 343 mmH2O, compared to 800 and 1700 to 3000 mmH2O of the DPF under the same operation conditions An electrostatic technique that generates unipolar ions and imposes a strong electrostatic force on the FTF could DEVELOPMENT OF ELECTROSTATIC DIESEL PARTICULATE MATTER FILTRATION SYSTEMS compensate for the low PM removal performance of the FTF while maintaining the required low pressure drop ACKNOWLEDGEMENT−This research was supported by a Basic Research Fund (SC0770) of the Korea Institute of Machinery and Materials REFERENCES An, S., Cho, G., Choi, H., Jeong, Y., Lee, E., Oh, K., Han, S., Kim, K and Park, S (2006) A study on the PM reduction of catalytic metal foam filter Fall Conf Proc., Korean Society of Automotive Engineers, Paper No 06F0055, 366−370 Bruck, R., Hirth, P., Reizig, M., Treiber, P and Breuer, J (2001) Metal supported flow-through particulate trap; A non-blocking solution SAE Paper No 011950 Cho, G., Choi, H., Jeong, Y., Kim, H., Ahn, S., Jeong, B., Choi, Y., Kim, D., Yoon, C S., Lee, E., Oh, K., Han, S., Kim, K., Park, S., Kim, G and Choi, S (2007) PM reduction performance and regeneration characteristics of catalyzed metal foam filters for a 3L diesel passenger vehicle SAE Paper No 013456 HEI (2002) Understanding the Health Effects of Components of the Particulate Matter Mix: Process and Next Steps HEI Perspectives, [Online] Health Effects Institute, Boston, MA, Available at http://www.healtheffects.org/ Pubs/Perspectives-2.pdf Hinds, W C (1999) Aerosol Technology 2nd Edn , 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Center for Diesel Research Report Majewski, W A (2008) Flow-through Filters, Diesel Net Technology Guide, [Online]Ecopoint Inc Available at http://www.dieselnet.com/tech/cat_ftf.html [Accessed 03 December 2008] Monaghan, M L (2000) Future gasoline and diesel enginesReview Int J Automotive Technology , , 1−8 Park, S J., Lee, D G., Kim, J., Cho, G., Kim, H and Jeong, Y (2007) Filtration characteristics of metal foam filters for DPF combined with electrostatic precipitation mechanism Trans Korean Society of Automotive Engineers , , 151−158 Park, Y., Choi, Y., Jung, H., Kim, N and Lee, J (2006) A study on the emission reduction performance of a partial flow diesel particulate filter Fall Conf Proc., Korean Society of Automotive Engineers, Paper No 06-F0037, 248−253 Yoon, C S and Cho, G (2009) Study of design and CFD analysis for partial DPF utilizing metal foam Trans Korean Society of Automotive Engineers , , 24−34 1 15 17 Copyright © 2010 KSAE 1229−9138/2010/053−02 International Journal of Automotive Technology, Vol 11, No 4, pp 455−460 (2010) DOI 10.1007/s12239−010−0056−7 EXPERIMENTAL INVESTIGATION OF THE VALVETRAIN FRICTION IN ACTUAL ENGINE OPERATION CONDITIONS S KANG , S K KAUH and K.-P HA 1)* 1) 2) School of Mechanical and Aerospace Engineering, Seoul National University, Seoul 151-744, Korea Power Train R&D Center, Hyundai Motor Company & Kia Motor Corporation, 772-1 Jangduk-dong, Hwaseong-si, Gyeonggi 445-706, Korea 1) 2) (Received 18 August 2009; Revised October 2009) ABSTRACT−Recently, the demands for improved fuel economy have been continually rising because of environmental protection policies, legislative pressures on emissions and increases in the price of oil Reducing the friction power loss in a production engine may be regarded as one of the most effective technologies for improving fuel economy because the technology is cost effective and applicable to a great number of vehicles This paper describes attempts to measure the torque needed to drive a camshaft and to examine the sources of the torque fluctuations in order to analyze the friction in valvetrains The measurements were performed through a cam sprocket-type torquemeter, which was able to measure the torque of the valvetrain under actual engine operating conditions In the cam torque measured, the fluctuations were mainly dependent on the primary oscillations caused by cam events and the secondary oscillations caused by the valvetrain natural frequency The range of the fluctuations became greater at high speed because of the inertial mass The resulting FMEP (friction mean effective pressure) of the valvetrain decreased, and the effective peak tension increased with an increase in the engine speed KEY WORDS : Torque, Valvetrain, Torquemeter, Friction, FMEP INTRODUCTION has been a tendency to increase engine speed and combustion pressure to improve the fuel efficiency and output Therefore, valvetrain components move with higher speeds and accelerations As a result, vibration, friction losses and engine noise have increased (Teodorescu et al., 2002) For this reason, testing and verification of valvetrains have become more important Most friction measurements on valvetrains were performed on a test rig in which the torquemeters were connected in-line between the driving motor and the actual cylinder head of the engine It was therefore very difficult to accurately measure the dynamic characteristics of a valvetrain in a real engine environment Crane and Meyer developed a comprehensive design tool that could be used to model an engine’s valvetrain friction components and that was validated by measuring the friction torque of the valvetrain and removing the valvetrain parts However, the test was made in a cylinder head, not a real engine A camshaft was motored indirectly using a pulley and a belt As a result, it is difficult to validate the model because of undesirable belt oscillations and tension changes (Crane and Meyer, 1990) Teodorescu et al isolated and determined the main components of the friction in a valvetrain system on a firing, single-cylinder diesel engine using strain gauges and an accelerometer (Teodorescu et al., 2002) Baniasad and Emes measured the driving torque of a valvetrain in a real engine using strain gauges and a slip ring However, each engine to be measured had to be For several decades, the efficient use of energy has been an important issue in all industrial fields Fuel economy in particular is one of the most important factors in evaluating overall vehicle performance in the automobile industry; therefore, fuel efficiency is considered one of the most important issues in this industry Recently, the demands for improved fuel economy have been continually rising because of environmental protection policies, legislative pressures on emissions and increases in the price of oil To improve fuel efficiency and satisfy exhaust regulations, several new technologies, including GDI (gasoline direct injection), VVT (variable valve timing) and cylinder deactivation, are being utilized Reduction of friction losses in a production engine can be regarded as one of the most effective technologies because the technology is cost effective and applicable to a great number of vehicles The valvetrain mechanism is one of the major focuses of engine development because it has an effect on the performance of spark ignition engines The friction losses related to the cam mechanism have become important as energy-conscious design becomes the new trend Although the valvetrain accounts for 6~10% of the total friction losses in an engine, attempts to reduce this number are being made (Gangopadhyay et al., 2004) Recently, there *Corresponding author e-mail: pigtiger.kang@gmail.com 455 456 S KANG, S K KAUH and K.-P HA modified to set up the slip ring (Baniasad and Emes, 1998) To compensate for this weak point, a torquemeter using telemetry was considered Recently, a cam sprocket-type torquemeter that can measure the torque of a valvetrain in a real engine has been developed The developed torquemeter has wireless communication via Bluetooth and a non-contacting power supply (Kang , 2007) The objective of the research discussed in this paper was to measure the torque of a valvetrain using a cam sprockettype torquemeter and to examine the complex dynamic mechanisms of valvetrains From the measurements, an investigation into the sources of the torque fluctuation was performed and the FMEP (friction mean effective pressure) of the valvetrain was determined et al INSTRUMENTATION 2.1 Cam Sprocket-type Torquemeter A camshaft is driven by a crankshaft; the two are connected to each other using a timing chain and sprockets for uniform valve timing The torque to drive the valvetrain is transferred through a cam sprocket A cam sprocket-type torquemeter can replace the existing cam sprocket and measure the torque using a torque sensor and strain gauges Figure shows a schematic view of the torque sensor The torque sensor is similar in shape to the shape of a cam sprocket and has four spokes to connect the rim to the hub The relationship between the strain generated in a spoke and the torque acting on the torque sensor can be derived from the superposition of two cantilevers through model simplification Considering the moments acting on the hub, at equilibrium, we have the following if the number of spokes is n: T - +M−Fr=0 n (1) The bending moment at the end of the spoke is R – r⎞ M=⎛⎝–T - ⋅ -n R + r⎠ (2) The strain on the spoke can be calculated from the following equation of the bending moment: Figure Schematic view of the torque sensor Figure Position of the installed strain gauge and a wiring diagram of a full bridge ε x My = - σ = - E (3) x x EI The strain on the spoke at = /4, which is where the strain gauge is located, can be determined as follows: x εx l x = l R–r T = ⋅ ⋅ - , Ebt R + r n (4) Equation (4) shows that the strain is linearly related to the torque T, which is transferred from the engine to the camshaft, and is dependent on the shape of the torque sensor and the point of installation The strain gauges were installed symmetrically at two spokes, as shown in Figure 2, and were wired to a Wheatstone full bridge for thermal stability (Kang , 2007) et al 2.2 Calibration Figure shows a comparative calibration apparatus for determining the relation between the output voltage of the torque sensor and the torque value applied to the torquemeter The cam sprocket-type torquemeter and the master torquemeter were connected in series, and the right side of the cam sprocket-type torquemeter was fixed, while the left side of the master torquemeter was connected to a cantilever to weigh the balance weights A bearing was installed to support the cantilever to reduce the bending moment caused by the mass The experiment repeatedly loaded and unloaded the mass, and the applied torque in the calibration experiment was set using the value from the master torquemeter The comparative calibration result is presented in Figure The cam sprocket-type torquemeter showed results with good linearity, and the Figure Comparative torque calibration apparatus 596 J HUR From the analysis result, it can be estimated that and by using the linear curve fitting method, and the estimated core loss resistance can be calculated from (4) and (5) Figure shows the estimated results of the core loss, eddy current resistance ( ) and hysteresis loss resistance ( ) for the prototype IPMSM Also, d and q are obtained directly by using the armature linkage flux and magnet linkage flux, which are obtained from the FEA Thus, d and q according to the armature current vector can be solved by using (6) (Nakamura , 2003; Lee , 2006) Rce Rch Rc Rce Rch L L L L et al et al ψo cos α – ψa Ld = -id ψo sin αLq = -iq (6) (5) where, ψo is the total linkage flux of the armature current reaction and α is the phase difference between ψa and ψo Figure shows the d and q of the prototype IPMSM, which varies depending on the - and -axis currents, respectively These calculated inductances are used in the torque calculation of the prototype IPMSM The constant torque characteristics of the prototype IPMSM are achieved using the nonlinear equivalent magnetic circuit method based on the current vector control theory and the parameters of the IPMSM as mentioned above The current vector is controlled in order to produce the maximum torque per current in the constant torque region Figure Inductance profile according to the - and -axis currents Figure 10 Cross section of the initially designed IPMSM for the EHPS system and the design variables for optimization Figure No load core loss ( ce) and equivalent core loss resistance, ( ch) for prototype IPMSM R R L 1 1- = + Rc Rce Rch × f (4) After calculating the total iron loss ( C) obtained from Finite Element Analysis (FEA) at each operating frequency (Hur, 2008), the iron loss resistance ( c) can be calculated as shown below in (5) W R V Rc = where, V = vod + voq Wc 2 d q L d q DEVELOPMENT OF AN ELECTRIC MOTOR-DRIVEN PUMP UNIT OPTIMIZATION OF THE IPMSM FOR IMPROVEMENT IN PERFORMANCE 597 Figure 11 Main effect of the cogging of each design parameter cogging torque, namely, chamfer, tooth surface, and effective open-angle of the permanent magnet in the rotor The three design variables are shown in Figure 10(b), (c) In this paper, the optimal shape design has been accomplished by using the response surface method (RSM) for MTPC control, which is well adapted to making an analytical model for a complex problem The RSM provides the designer with an overall perspective of the system’s response to the behavior of design variables within a design space(Fernandez-Bernal ; Kim , 2006) The effect of the cogging torque for each design variable is shown in Figure 11 This result is obtained using response surface analysis The cogging torque of the IPMSM decreases as the chamfer and the effective open-angle of the permanent magnet increase Figure 12(a) shows the result of the comparison between the initial and optimized model Figure 12(b) shows measured results about the optimal condition Although, the experimental result is a little larger than the analysis result and the measured wave is more distorted than that obtained through analysis, the required level of the cogging torque can still be obtained by using the response surface method Figure 13 shows the torque characteristics of the torque commands for the initial model and the optimized model Below the base speed, the maximum torque is produced by Figure 12 Characteristics of cogging torque for the designed model Figure 13 Torque characteristics for the developed IPMSM Figure 10(a) shows the configuration of the prototype IPMSM designed using parameter verification and characteristic analysis from the nonlinear equivalent magnetic circuit model The optimization is performed for the improvement of the torque characteristics, including the reduction of the cogging torque Based on the initial design of the prototype motor, three variables have been considered for the improvement of the torque characteristics including the reduction of the et al et al 598 J HUR Figure 16 Configuration of the IPMSM with experimental devices Figure 14 Results of armature current calculation for command torques Figure 17 Comparison of experimental results between initial model and optimized model Figure 15 Efficiency distribution for the overall operation point the maximum torque per ampere control Both models satisfy the requirements for torque in EHPS to operate the pump Figure 14 shows the results of armature current analysis concerned with torque commands The armature current is constant at the rated torque command, but at torque values smaller than the rated torque, the armature current increases with higher speeds as the voltage limit is reached Figure 18 Bench test set for the electric motor-driven pump unit DEVELOPMENT OF AN ELECTRIC MOTOR-DRIVEN PUMP UNIT 599 Figure 22 Developed all-in-one type EHPS system Figure 19 Experimental results of the actual hydraulic pressure the initial model in order to generate the same torque It appears that the electric power consumption of the optimized model is relatively smaller compared to the initial model and hence, the optimized model has a higher efficiency at whole driving points EXPERIMENTAL RESULTS Figure 20 Current wave of the relief condition at 4,000 rpm Figure 21 Experimental power characteristics of the electric motor- driven pump unit Figure 15 shows the efficiency map of whole driving points of the initially designed model Here, in order to solve the efficiency, the copper loss and the core loss have been included, but mechanical loss has been ignored From this result, the optimized model requires less current than Figure 16 shows the experimental devices for the IPMSM The experimental results of the initially designed model and the optimized model are shown in Figure 17 From the results, it can be seen that the performance of the optimized model has improved with respect to the initial model The bench test set for the electric motor-driven pump unit of the EHPS system is shown in Figure 18, and the test results of the comparison between the fluid quantity and the hydraulic pressure are shown in Figure 19 At the no-load condition, the actual hydraulic pressure is biased from [kg/cm2] to [kg/cm2] according to the speed; although the maximum relief pressure required is 120 bar, the hydraulic pressure was measured and determined to be 95 kg/cm2 Figure 20 shows the current waveform at the relief condition and at 4000 rpm and the applied well vector control as an approximate sinusoidal wave The power characteristics the electric motor-driven pump unit are described in Figure 21 The maximum efficiency of the electric motor-driven pump unit is about 40%, and occurs between 30 kg/cm2 and 60 kg/cm2 of the hydraulic pressure Figure 22 displays the all-in-one type electric motor-driven pump consisting of the ECU, the IPMSM, and the vane-type hydraulic pump CONCLUSION This paper presents the development of the electric motordriven pump unit for the EHPS system using a 42V powerNet In order to improve the system, the use of the IPMSM as a pump motor has been investigated using the equivalent magnetic circuit method, taking iron loss into consideration for analysis and design Also, shape optimization has been performed for the improvement of torque characteristics and electric power consumption 600 J HUR Performances of both the pump unit and the IPMSM have been verified by experimental results Finally, an allin-one type of an electric motor-driven pump unit has been developed after experimental verification ACKNOWLEDGEMENT−This work was supported by the Human Resources Development of the Korea Institute of Energy Technology Evaluation and Planning(KETEP) grant funded by the Korea government Ministry of Knowledge Economy (No 2008-E-AP-HM-P-10-0000) REFERENCES Bianchi, N., Dai Pre, M and Bolognani, S (2006) Design of a fault-tolerant IPM motor for electric power steering IEEE Trans Vehicular Technology 55, 4, 1102−1111 Eki, H., Teratani, T and Iwasaki, T (2007) Power consumption and conversion of eps systems Power Conversion Conf 2007 (PCC07), 1333−1339 Fernandez-Bernal, F., Garcia-Cerrada, A and Faure, R (2000) Model-based loss minimization for DC and AC vector-controlled motors including core saturation IEEE Trans Industry Applications 36, 3, 755−763 Hur, J (2008) Characteristic analysis of interior permanentmagnet synchronous motor in electrohydraulic power steering systems IEEE Trans Industrial Electronics 55, 6, 2316−2323 Kim, S I., Bhan, J H., Hong, J P and Lim, K C (2006) Optimization technique for improving torque performance of concentrated winding interior PM synchronous motor with wide speed range Proc IEEE IAS Annual Meeting, 4, 1933−1940 Lee, J Y., Lee, S H., Lee, G H., Hong, J P and Hur, J (2006) Determination of parameters considering magnetic nonlinearity in an interior permanent magnet synchronous motor IEEE Trans Magn 42, 4, 1303−1306 Nakamura, K., Saito, K and Ichinokura, O (2003) Dynamic analysis of interior permanent magnet motor based on a magnetic circuit model IEEE Trans Magn 39, 5, 3250− 3252 Sebastian, T., Islam, M S and Mir, S (2005) Application of permanent magnet synchronous machines in automotive steering systems KIEE Trans Electrical Machinery and Energy Conversion Systems 5-B, 2, 111−117 International Journal of Automotive Technology, Vol 11, No 4, pp 601−610 (2010) DOI 10.1007/s12239−010−0072−7 Copyright © 2010 KSAE 1229−9138/2010/053−18 MODELING, PARAMETER ESTIMATION AND NONLINEAR CONTROL OF AUTOMOTIVE ELECTRONIC THROTTLE USING A RAPID-CONTROL PROTOTYPING TECHNIQUE R GREPL and B LEE 1)* 2) Mechatronics Laboratory, Faculty of Mechanical Engineering, Brno University of Technology, 616 69 Technicka 2, Brno, Czech Republic 2) Department of Mechanical and Automotive Engineering, Keimyung University, Daegu 704-701, Korea 1) (Received 23 January 2009; Revised 23 November 2009) ABSTRACT−An electronic throttle consists of a DC motor, spur gears, a return spring, a position sensor, power electronics and an electronic control unit Fast and precise position control of this electromechanical system is relatively difficult due to very high friction and the strong nonlinearity of the spring Simple application of linear control, such as PID, fails In this paper, two new controller structures suitable for different reference signal types are described The key component of the position controller is the friction compensator based on either/both feedforward or feedback principles The quality of the resulting behavior was measured using several criteria including the measure of control activity around the equilibrium position The control activity directly influences the vibration, the noise and the wear of the servo system The proposed controllers demonstrated superior behavior compared with other published structures KEY WORDS : Electronic throttle control, Nonlinear control, Friction compensation, Parameter estimation, Rapid control prototyping NOMENCLATURE ϕ ϕref ϕLH u e G Ts N kP kI Td i kemf L me mM τ R i12 η12 kS bemf b : angle of throttle shaft : reference (desired) value of angle : LH position of throttle plate : control input to throttle, voltage normalized to interval (−1;1) : position (angle) error : gain limit of derivation : computation sample time : number of sample times : proportional gain in PID : integral gain in PID : derivative time constant in PID : electrical current : constant of back electromotive force : armature inductance : electrical torque recalculated to throttle shaft : electrical torque on motor shaft : mechanical torque : armature resistance : total gear ration from motor to throttle shaft : total efficiency of mechanical system : stiffness of the return spring : mechanical viscous friction Jred J kxy N µ p z n d λ α TS : damping of the electromechanical system, which applies to the mechanical viscous damping and the back electromotive force : mechanical inertial moment reduced to the valve shaft (includes inertia of the DC motor, gearbox, and plate) : inertial moment in normalized units : stiffness of nonlinear spring in given part xy : normal force : dry friction coefficient : state value in Reset Integrator model : state value in LuGre model : nr of samples for error measurement : value of dead zone of the friction compensator : parameter of SMC : parameter of SMC : sample time INTRODUCTION One important current area of research and development in the automotive industry is focused on the X-by-wire concept The main idea is adopted from aircraft, where “flyby-wire” is the standard approach used in military as well as civil aircraft Computer control has replaced conventional mechanical controls; the actuator is connected to the pilot only by means of a “wire” As a result, the maneu- *Corresponding author e-mail: grepl@fme.vutbr.cz 601 602 R GREPL and B LEE verability of the aircraft increases significantly due to computer control of a naturally unstable design Fly-bywire is very often mentioned as a typical mechatronic solution In automotive design, there are many variants of this philosophy in use already or in development Examples include steering-by-wire, where conventional mechanical steering is replaced by a sensor and an electrical servomotor, brake-by-wire, where conventional hydraulic or pneumatic actuators are replaced by electrical ones, and (most commonly) throttle-by-wire, where a sensor and an electrical motor replaces a mechanical linkage between the accelerator pedal and the throttle (Stence, 2006) The throttle-by-wire system consists of a pedal sensor, the throttle body (with a DC motor, spur gears and a potentiometer as the position sensor) and an electronic control unit (ECU) Thus, the conventional mechanical linkage of pedal and throttle via a bowden cable is replaced by a mechatronic design Full ECU control of throttle plate behavior enables better fuel economy and emissions, and it provides the possibility of using advanced tracking control algorithms or other overall system improvements Due to mass production of automotive parts and the corresponding compromises between technical quality and manufacturing costs, high friction is a problem that plagues the throttle mechanism Moreover, safety regulations require the return of the valve to a slightly open position–a socalled “limp home” (LH) mode–in case of system failure This feature is implemented by a relatively strong spring near the LH position However, using this spring stiffness through the full range of the valve would also produce an enormous motor load along with significant energy consumption and heating In most throttles, a nonlinear spring is used to solve this problem (Pavkovi c , 2006, Deur , 2003) Here, we describe a solution to this problem by illustrating how the advanced control algorithm implemented in the ECU can significantly improve inexpensive electromechanical systems In the last decade, many authors have published interest′ et al et al ing results using electronic throttle control; many of them have been directly related to automotive firms (Pavkovi c , 2006) The high nonlinearity of the system disqualifies the use of simple linear controls (e.g., PID) It is well known that linear controllers cannot deal with the dry friction phenomenon because steady-state error arises in the case of a PD regulator and oscillations arise in the case of a PID controller The multi-model approach for throttle modeling has been successfully studied previously (Hadilebbal , 2007) In this approach, a nonlinear model is replaced by several linear models, and the algorithm switches between them The work is highly inspired by the PWARX (Piece Wise Auto Regressive eXogenous) method Other work (Trebi-Ollennu and Dolan, 2004) uses an adaptive fuzzycontrol approach applied on unmanned ground vehicles The whole drive system is modeled and controlled with the goal of low-speed control, and a detailed model of the throttle control is not provided Contreras have identified throttle parameters (Contreras , 2002) In their approach, the dynamic friction model of LuGre is used with different inertial reduced moments of mechanical throttles, which are considered in terms of the closing or the opening direction This model does not discuss its control scheme One highly inspiring set of papers has been published by Pavkovic , Deur, and Vasak (Pavkovic , 2006; Deur , 2003; Vašak , 2007) In these works, the controller architecture consists mainly of a friction compensator, a LH compensator, and a PID element There are also other improvements including variable filtering according to the control error, feedforward elements, and gain scheduling of the PID At a given moment, only one of the friction or LH compensators is used Deur (Deur , 2003) describes the self-tuning of compensator parameters Baotic, a collaborator of Deur, (Baotic , 2003) uses Model Predictive Control and a Karnopp friction model The key issue in effective throttle control is compensation for friction Because friction is a nonlinear function of velocity, several authors deal with the compensators based on measured or reconstructed velocity (Olsson , 1998) The authors note that the approach is dependent on the precision of the velocity measurement or estimation A much better alternative to this approach is the use of position error as the input for the friction compensator (Iserman, 1996; Yang, 2004; Pavkovi c , 2006; Deur , 2003) Sliding Mode Control (SMC) for ETC is also an active area of research Beghi (Beghi , 2006) and Zhang (2006) use SMC in conjunction with a Sliding Mode Observer and compare it with a Kalman filter Also, many researchers use the throttle model as a part of other complex models including the engine, the transmission, the wheels and tires, and the traction control system (Ryu , 2005; Jung , 2000; Ishikawa , 2007) ′ et al et al et al et al ′ al ′ et al et et al et al et al et al ′ et al et al et al Figure Photo of an electronic throttle body et al et al et al et al MODELING, PARAMETER ESTIMATION AND NONLINEAR CONTROL OF AUTOMOTIVE ELECTRONIC This paper outlines the authors’ development and extension of these cited works in the following ways: • Development and testing of two new controllers suitable for smooth and staircase signals Additionally, a feedforward compensation for nonlinear springs was developed, used, and tested; and it proved superior to other published feedback compensation methods • Use of multiple-criteria comparisons of different control structures including PID; all controllers were tested using two sample times (minimal and realistic) • Description of an optimal approach for parameter estimation for the system based on closed-loop data acquisition According to other authors, e.g., (Baric , 2005) and (Pavkovi c , 2006), the air flow over the throttle valve was not included in the experiments and was considered as a disturbance The Rapid Control Prototyping (RCP) technique was used with dSpace hardware for maximal time efficiency in the development process Preliminary experiments and results which have been further developed and presented in this paper was published by authors in (Grepl and Lee, 2008a) and (Grepl and Lee, 2008b) et al ′ MODEL OF THE ELECTROMECHANICAL THROTTLE 2.1 Mechanism and DC Motor The electromechanical part of the ETC consists of a brushed DC motor, spur gears, a nonlinear return spring, a throttle plate, and a potentiometer (Figure 2) The mechanical equation of motion is of the following form: (1) Jredϕ·· =me – τviscous – τspring – τfriction The expression of electrical equilibrium is the following (using the standard linear model of DC motor): u=Ri+L d i +kemfϕ· M dt mM=kemf i (2) 603 of the DC motor, given by inductance In this case, the model is static and takes the following form: L u=Ri +kemfi12ϕ kemfi12- · i= -1 u− -ϕ R R (5) The expression for electrical torque is the following: 2 i η kemfi12η12- k u− emf 12 12- ϕ· me= R R (6) The viscous mechanical torque is of the following form: (7) τviscous =bmechϕ· Next, Equation can be rewritten as the following: 2 i η kemfi12η12- k Jredϕ·· = u− emf 12 12- ϕ· −bmechϕ· R R − τspring − τfriction (8) k2emfi122η12-⎞ ϕ· kemfi12η12- ⎛ Jredϕ·· = u−⎝ bmech + R R ⎠ − τspring − τfriction (9) Equation (9) can be divided by kemfi12η12/R to obtain the model in units of voltage: (10) Jϕ·· =u− bϕ· − uspring ( ϕ )−ufriction( ϕ· ) The spring torque applied to the shaft can be generally expressed in the form: (11) τspring =kS( ϕ ) ϕ The stiffness, S, varies with respect to the throttle position (Figure 3) According to Equation (10), the spring torque in units of voltage, spring(ϕ), is used The stiffness, HS, expresses the hardstop model stiffness if it is needed to simulate the limitation of plate movement in the model For normal throttle operation (within limits), the static spring equation is of the following form: k u k (3) Gears introduce the following relations between the motor and plate variables: (4) ϕM=ϕi12 me =mMi12η12 Usually, one can neglect the dynamics of the electrical part Figure Schematic of the electronic throttle Figure Schema of the nonlinear spring (LH – limp home; CL – throttle closed; OP – throttle opened; HS – the stiffness of hardstops, if modeled) 604 R GREPL and B LEE ⎧ k ( ϕ – ϕ ) , –u < u < u ⎪ uS =⎨ u + k ( ϕ – ϕ ) , u > u ⎪ –u + k ( ϕ – ϕ ) , u > u ⎩ LH LH LHO LHC OP LHC LH CL S S LH 1998) This model is of the following form: LHO (12) LHO S LHO z·=v−σ v - z g(v) v ⎛ ⎞ – - (15) ⎝v ⎠ 2.2 Friction Models Usually, the very complex phenomena of friction is considered and modeled as a combination of viscous and dry friction The viscous component is simple to express, and according to Equation (9), it can be incorporated into the general damping of the electromechanical system, Dry friction in the throttle body system is relatively high due to the mass production and consequent low-cost design of the system The significance of friction is illustrated in Figure Friction accounts for 10% of the input voltage range The impact of friction on control is even more significant because of the return spring’s nature The most simple Coulomb friction model, (13) (ϕ)µN τ (ϕ) has several well known limitations: a) friction is not defined for zero velocity; b) the model does not cover the increase of friction at near-zero velocities; and c) a nonsmooth system is problematic for simulation The literature details numerous improvements to the static model, such as including the Stribeck effect, but static friction models are often not sophisticated enough for effective simulation For this reason, two dynamic friction models are described b C · · = –sgn , 2.2.1 Reset Integrator model This dynamic friction model effectively describes the stiction effect at near-zero velocities A new dynamic state, , is considered and can be understood as the bending of a virtual bristle Proper behavior requires a zero detection feature: p ⎧0 p· =⎨ if ⎩v F fric = ( v > ∧ p ≥ p ) ∨ ( v < ∧ p ≤ –p ) 0 g( v) =α +α e F =σ z +σ z· +( α v ) 0 fric 1 Thus, it is necessary to estimate five parameters: α , α are kinetic friction and increase of stiction, respectively (see and in the Reset Integrator model for comparison); is the Stribeck velocity; σ , σ are the stiffness and damping of “bristles,” respectively; and α is the normal viscous friction (already included in Equation (9)) Both dynamic friction models were used in the parameter estimation described in Section The LuGre model better captures system properties, whereas the Reset Integrator is more effective for simulation This feature can be very important in cases when the throttle model is only one part of a complex model consisting of an engine, a transmission, and/or vehicle dynamics a Fkin v0 PARAMETER ESTIMATION OF NONLINEAR MODEL An important class of advanced nonlinear control algorithms (especially friction compensators) requires modeling of the plant (structure, equations) and good estimates of its parameters This section describes and the search for the parameters (PE, Parameter Estimation) of the aforementioned model 3.1 Initial Experiment Using Open Loop Response The most important information about the system properties can be obtained using the response to a slow sinusoidal input, Figure shows the measured current and angle, ϕ, in the time domain Figure shows the ϕ− plot This experiment confirms the following: a) the stiff nature of the return spring near the LH position, b) the significant fricu u otherwise (1 + a(p))F kin p + p ⎧a a(p)=⎨ ⎩ (14) β p· if p d e + d, if e < d 0, else The dead zone used in the friction compensator significantly reduces chattering around the zero error position and thus improves controller properties Controller B – With this controller, PID and spring FFC are used as in variant A The friction compensator is implemented by using the Smooth-Sliding Mode Control (SSMC) (Figure 10): u=uPID+uspring+uSSMC uspring =f ( ϕref – ϕLH ) uSSMC=uF tanh( α( λe + e· ) ) (24) A detailed description of SSMC and the settings of α and λ can be found in (Young , 1999) Controller C – In this variant, the PID and spring FFC are implemented as in variant A The friction compensator is implemented using SSMC combined with a simple positionerror friction compensator that is switched off at every odd sample time The compensator is shown in Figure 11 and is of following form: et al u=uPID+uspring+uC ⎧ ⎪ uC = ⎨ ⎪ ⎩ uSSMC, e > e u sign( e )p( t ) 0, e < e S F (25) p where S and are the switching values obtained experimentally and ( ) is a rectangular pulse signal with a period e ep p t MODELING, PARAMETER ESTIMATION AND NONLINEAR CONTROL OF AUTOMOTIVE ELECTRONIC 607 of 2TS and a duty factor 0,5 Here, the PID and spring FFC are implemented as in variant A The friction compensator is identical to that in variant A (23) but is implemented as a feedforward For the computation of discrete derivation is used Equation (22) Controller D – 4.3 Implementation Details 4.3.1 Testing reference signal Two types of position reference angle were used in our experiments: 1) a continuous reference signal that is a smooth signal simulates the slow change of required throttle position, and 2) a staircase reference signal that simulates the sudden change of plate angle Both signals, shown in Figure 13, were generated randomly (except for the first part of the continuous signal) These signals were then stored and used in all experiments Randomly generated signal is a good representation of variable driving conditions and covers virtually all possible changes of the reference Figure 11 Block diagram of friction compensation using off-pulses near the zero error (Controller C) 4.3.2 Sample time and modeling of computational delay Every practical design of ETC must address the limited computational power of low-cost microcontrollers used in cars For each sample time, one must read the position of the throttle plate (i.e., ADC), compute an action using a given controller and set outputs (i.e., PWM, digital IO, or, for our case, DAC) The total time of these procedures, plus communication with the upper level of control, determines the minimal sample time Figure 14 shows the timing schema on the microcontroller and the model used on Figure 12 Block diagram of control with friction feedforward compensation (Controller D) Figure Block diagram of control with friction feedback compensation using position error (Controller A) Figure 13 Continuous (smooth) and staircase reference angle signal used in all control experiments (see Table and Table 3) Figure 10 Block diagram of control with friction feedback compensation using SMC (Controller B) Figure 14 Timing on a real microcontroller (top) and a simulation of a microcontroller computational delay using RCP hardware from dSpace (bottom) 608 R GREPL and B LEE Figure 15 Schema of the RCP with dSpace modular hardware the RCP hardware On the RCP hardware, computational time is very small The goal is to simulate a delay similar to those seen in reality 4.3.3 Hardware setup The RCP-hardware used in our experimental setup is based on modular dSpace technology It consists of a main processor (DS 1005 PPC), a board (DS 2103) with 14-bit D/A converters, a board (DS 2003) with 16-bit A/D converters, and communication cards (DS 814 and 815) RealTime Workshop in Simulink generates C-code for the DS 1005 PPC target, and the RTI (Real-Time Interface) provides a link between the code and the dSpace hardware With a goal to minimize the hardware delay in the control loop, the operational power amplifier (SR4409439) from aPEX microTEX was used for the DC motor actuation The physical measured voltage on the throttle potentiometer in the range of 0.25~2.25 V (2.25 V corresponds to full open) was used as the unit in this work The input voltage, was normalized in the range (−1;1) Figure 16 Responses using the continuous (smooth) reference Sample time, 50us; PID controller (top) vs Controller D u, 4.1 Experimental Results The controller structures described in Section 4.2 were tested using the two reference signals defined in Sec 4.3.1 Table Multi-criteria evaluation of the controllers’ behavior–continuous (smooth) reference signal Minimal sample time 50us Realistic sample time 5ms MSE MAE maxE CoEf MSE MAE maxE CoEf 1e-3 1e-3 1e-3 1e-3 1e-3 1e-3 PID 0,103 4,18 0,175 75,5 0,698 14,2 0,252 101 A 0,108 3,43 0,165 84,4 1,08 13,5 0,308 66,6 B 0,404 6,76 0,212 65,5 0,878 13,6 0,272 75,3 C 0,119 4,28 0,168 80,6 0,902 13,7 0,292 74,5 D 0,119 4,30 0,169 44,6 0,835 12,6 0,290 43,1 Figure 17 Responses using the continuous (smooth) reference Sample time, ms; PID (top) vs Controller D Tables and summarize the results evaluated via the criteria introduced in Section 4.1 Figures 16~19 show the comparison of the system’s response using PID and the MODELING, PARAMETER ESTIMATION AND NONLINEAR CONTROL OF AUTOMOTIVE ELECTRONIC 609 Table Multi-criteria evaluation of the controllers’ behavior–staircase reference signal Minimal sample time 50us Realistic sample time 5ms MSE MAE maxE CoEf MSE MAE maxE CoEf 1e-3 1e-3 1e-3 1e-3 1e-3 1e-3 PID 0,414 14,6 0,0491 34,1 0,850 20,4 0,0849 40,4 A 0,00145 0,968 0,0049 79,5 0,0499 4,80 0,0299 20,2 B 0,00116 0,863 0,0039 70,4 0,0158 2,80 0,0150 44,4 C 0,00183 1,1 0,0053 0,660 0,00214 1,04 0,0089 15,5 D not working well not working well Figure 19 Responses using the staircase reference Sample time, 5ms; PID (top) vs Controller C Figure 18 Responses using the staircase reference Sample time, 50us; PID (top) vs Controller C best candidate controller Only two seconds (9.5~11.5 sec) are shown in detail, and the criteria in the Tables cover all experimental data (0~20 sec.) A system with a smooth reference signal can be controlled using PID only, but the proportional gain is very high and can cause significant chattering (noise) around the target position (Figure 16) Controller D notably reduces the control effort (approx 50%) given by CoEf and having similar MSE, MAE and maxE values This result applies for both sample testing times (50 us vs ms) The staircase reference signal disqualifies the PID from consideration, primarily in the realistic sample time of 5ms This well-known disadvantage of using PID to control systems with high dry friction is fully visible here Figure 19 (left) shows unacceptable steady state error Also, Controller D does not work well in this case The staircase signal has either an infinite- or zero-derivative; thus, friction compensation in feedforward control cannot be used Conversely, Controller C proves superior for control of the system Table summarizes the very good results given by this system, especially for the realistic sample time of ms The new controller feature, which switches off the action every even sample time near the target, significantly reduces the overcompensation and consequent oscillation around the reference CONCLUSION Two new controller structures have been presented in this paper The quality of the controllers has been measured using standard mean square error and by the consideration of required control effort, which has a direct influence on noise, vibration and wear of the throttle servo system Compared with reports in previous publications, feedforward compensation for spring nonlinearity was used and has been verified as significantly more efficient For the smooth reference signal, feedforward friction compensation is proposed along with the use of a derivative impulse-area invariant For the staircase reference signal, an improved smooth sliding-mode controller was developed that switches off its activity every odd sample time, thus leading to an important decrease in control error and 610 R GREPL and B LEE control effort For practical applications, the implementation of a hybrid structure is recommended In this case one of the two best candidates is switched on according to the derivative of the reference signal Two main steps for effective parameter estimation are also described First, an open-loop quasistatic input signal is used to obtain the basic characteristics of the system Then, a closed-loop scheme is used for the measurement of the system’s dynamic response Many experimental sets have been prepared, and the Nelder-Mead simplex search has been employed for offline parameter estimation All practical experiments were carried out using the Rapid Control Prototyping modular dSpace hardware ACKNOWLEDGEMENT−This work has been supported by the Automotive Mechatronics Parts Nurturing Group at Keimyung University, by the Ministry of Knowledge Economy (MKE) and Korea Institute of Industrial Technology Evaluation and Planning (ITEP) through the Center for 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