MEASUREMENT OF FRICTION IN INTERNAL COMBUSTION ENGINE

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1 XLI. INTERNATIONAL SCIENTIFIC CONFERENCE OF CZECH AND SLOVAK UNIVERSITY DEPARTMENTS AND INSTITUTIONS DEALING WITH THE RESEARCH OF COMBUSTION ENGINES SEPTEMBER 6-7, 2010 LIBEREC, CZECH REPUBLIC TECHNICAL UNIVERSITY OF LIBEREC DEPARTMENT OF VEHICLES AND ENGINES MEASUREMENT OF FRICTION IN INTERNAL COMBUSTION ENGINE Emrich Miloslav 1, Fuente David 2, Rudolf Milan 3 Abstract This article describes a process and results from measurements of mechanical friction in internal combustion engine as affected by oil temperature. Oil and cooling water temperatures were controlled in a wide range. Spark plug adaptor with integrated pressure transducer was used for IMEP determination. Obtained results were compared with engine motoring method. DC dynamometer was used for breaking and motoring the engine. Insufficient accuracy of torque measurement with reversible weight mechanism led to installation of measurement reaction force of dynamometer stator by strain gauge. This in-house designed arrangement is also described here. 1. INTRODUCTION The difference between indicated and effective power is called a friction power. The friction power reduction is relevant to a fuel consumption reduction, which in turn influences a CO 2 reduction. A mechanical efficiency of internal combustion engine (ICE) is 0% at IDLE and about 90% at high operating load [1]. There are several methods for engine friction estimation. Two of them are currently being applied the most frequently. The first method is based on a break mean effective pressure (BMEP) and an indicated mean effective pressure (IMEP) measurement using an in-cylinder pressure indication on firing engine. This method gives the most accurate results. The second method is useful for determination of a partial contribution of friction couples. It is measured without any combustion, the engine is only motored and the in-cylinder pressure indication is used for the pumping loss evaluation. 1 Emrich Miloslav, Czech Technical University in Prague, Faculty of Mechanical Engineering, Department of Automobiles, Internal Combustion Engines and Railway Vehicles, Technická 4, Praha 6, miloslav.emrich@fs.cvut.cz 2 Fuente David, Czech Technical University in Prague, Faculty of Mechanical Engineering, Department of Automobiles, Internal Combustion Engines and Railway Vehicles, Technická 4, Praha 6, david.fuente@fs.cvut.cz 3 Rudolf Milan, Czech Technical University in Prague, Faculty of Mechanical Engineering, Technická 4, Praha 6, milan.rudolf@fs.cvut.cz

2 Main goal of presented work was to determine friction power of combustion engine at different oil and cooling water temperatures. Experimentally acquired data were required for calibration of a mathematical model described in another KOKA 2010 article. Measuring equipment and procedures are described here including the conditioning set-up. 2. MEASUREMENT OF FRICTION Experiments were performed in the combustion engine laboratory of author s department. The main equations for calculation of FMEP and mechanical efficiency η are: FMEP= IMEP BMEP (1) BMEP η = (2) IMEP BMEP 4 π M V t =, (3) where Mt is engine torque and V is engine displacement. The most accurate method to calculate IMEP is to indicate pressure in each cylinder. IMEP is then calculated as average value - see (4). IMEP p dv = 1 V #cylinders #cylinders i cycle 1 =, (4) where p is indicated pressure, V 1 is displacement of one cylinder, #cylinders is number of cylinders. In case of indication in only one cylinder, IMEP is calculated as (5) and it is assumed that indicated power is the same in all cylinders of the engine. p dv IMEP = (5) V cycle 1 This assumption was validated using evaluation of unevenness of instantaneous crankshaft speed. There were 125 teeth on the fly-wheel. Hall sensor was used to register these teeth. Signal from this sensor was connected to an in-house made counter card with base tick frequency of 50 MHz. Measured time between two teeth was used to calculate instantaneous RPM. This curve of actual RPM was decomposed using FFT and than first 200 harmonic frequencies were composed back to get smooth curve. Figure 1 shows two examples of actual RPM. Same high of RPM peaks enables subjective assessment of uniformity of indicated power in each cylinder.

3 2600 Measured RPM 3600 Measured RPM 2580 FFT (200 harmonic) 3580 FFT (200 harmonic) RPM 2500 RPM Crank angle Crank angle Figure 1: Speed irregularity at 2500 min -1 full load (left) and 3500 min -1 part load (right) 2.1 Engine and test-bed set-up Test-bed is equipped with DC dynamometer MEZ 1DS 736 V (rated power 90 kw) with ability to drive or break the engine at constant rotation speed. Three cylinder gasoline ICE Skoda 1.2 HTP (44 kw) was chosen for experiments Cooling system Figure 2 describes a modified cooling system of the engine. An original cooling system of the engine consists of two circuits. 1 Heater 6 kw 2 Electric water pump 3 Engine water pump 4 Expansion tank 5 Thermostat-removed 6 Tank external cooling water flows through 7 Heat Exchanger 8 Solenoid valve Thermocouples: two, twi, tw2, tw3 Engine temperature sensor: R2c Figure 2: Cooling water circuit The cooling water is circulating through the first circuit all the time the engine is running. This circuit contains a heat exchanger with a fan for passenger compartment

4 heating. The second main one circuit is normally controlled by thermostat. The big heat exchanger, which is usually placed in front of the car, is a part of this circuit. Certain modifications have been performed. The heat exchanger in a small circuit has been replaced by a resistance heater (6 kw) and an electric water pump Grunfos. The thermostat has been removed for this measurement. Four thermocouples (J type) have been installed into the both circuits. Cooling water mass flow rate has been controlled by solenoid valve (position 8 at Figure 2). This solenoid valve has been used for opening/closing intake pipe of a tank. It can be switched manually or controlled automatically depending on the temperature measured by a thermocouple denoted twi. The heater is switched by contactor controlled from a computer. Electric water pump is always on for safety. Described set-up with electric pump and without thermostat has many advantages. It is possible to preheat or cool down the cooling water on stopped engine before measurement. Temperature control using the solenoid valve, enables that cooling water temperature can be held within narrow band of ±2 C Oil conditioning system Figure 3 describes an oil conditioning system consisting of these main parts: centrifugal pump, two resistance heaters, counterflow plate heat exchanger and ball valves. Main idea was to suck the oil from an oil tank (position 8) and return the same amount (position 9) to keep constant level of the oil in the tank. This set-up does not influence lubricating circuit of the engine. Baroscope in the cylinder head was replaced by a pressure sensor. Reference oil temperature (toil) is measured in an oil tank. Figure 3: Oil conditioning system 1 Resistance Heating (each 1250W) 2 Ball valve - inlet heaters (oil) 3 Electromotor for oil pump 4 Oil centrifugal pump 5 Ball valve inlet heaters (oil) 6 Ball valve heat exchanger inlet (cooling water) 7 Cooling water outlet 8 Sucking orifice in the bottom of an oil tank 9 Return oil pipe to oil tank 10 Oil pump outlet 11 Oil/Water heat exchanger 12 Oil pressure sensor 13 Engine Skoda 1.2 HTP Thermocouples: toil, toil1, toil2, toil3

5 Oil pump and heaters are turned on/off by contactor controlled from a computer. It is possible to preheat or cool down oil on stopped engine. Cooling or heating or the both regimes together are set manually by adjustment ball valves position 2 and 4 on figure 3. Heating power of resistance heaters was chosen with regard to carbonization of the oil. Specific heating power is 10 W/cm 2, total power is 2x1250 W. 2.2 In-cylinder pressure measurement In-cylinder pressure measurement is performed to get IMEP at running engine. During motoring of the engine, indication is used to establish pumping loses. As the measuring equipment is quite expensive, only one cylinder was indicated. Ignition spark plug in the first cylinder was replaced by the indication spark plug KISTLER 6117BFD47. Signal was processed by AVL Piezo Amplifier A03 and converted to an analog signal 0-10V. An incremental counter AVL Angle Encoder 365CC was connected to a crankshaft pulley. Analog and digital outputs were connected to an acquisition card NI PCI An in-house software Iti-onl [2] was used as the indicating software. Always 100 consecutive cycles have been recorded in each operating point. Evaluation has been done using an in-house software Intec v.21 [3]. 2.3 TDC determination TDC determination has big influence on mechanical efficiency calculation. One degree error in TDC determination causes error in calculation of mechanical efficiency approximately 3%. TDC has been set by evaluation of pressure records on motored and running engine. At first, data were acquired from motored engine with full open throttle. Under the assumption that the ratio of specific heats κ is constant, and that there is neither heat transfer nor crevice flows, the cylinder pressure can be determined from (5). κ p( α 0 ) V( α 0 ) p( α) = (5) V α ( ) κ Under these assumptions, the maximum cylinder pressure will occur at volume minimum. Therefore, TDC position has been moved using software shift of pressure record to the position of the peak of pressure as first shot. This condition bounds searched TDC position from one side. Further shift of the TDC to negative values moves peak of the pressure to expansion which is nonsense. Figure 4 shows influence of TDC shift to negative values with incorrect loop creation between compression and expansion line close to TDC at compression and expansion stroke. Figure 4: TDC shift influence for motoring engine at 1500 min -1 (logarithmic axes)

6 Note to TDC position: shift to negative values cause higher IMEP and lower mechanical efficiency and vice versa. Shift to negative values is limited by formation of the loop at motoring (see Figure 4). Shift to positive direction is limited by unusually high mechanical efficiency at full load. More accurate setting of TDC is usually done using a thermodynamic evaluation. Some thermodynamics method has been published e.g. [4]. Thermodynamic evaluation of in-cylinder pressure at running engine at various RPM and load has been performed to correct TDC position. The curve of heat release has been checked at many points. 2.4 Measurement procedure Cost effective set-up has been used for water and oil conditioning. It does have neither enough cooling/heating power nor sophisticated control to hold set temperatures in narrow band except operation points with cooling water at normal operating temperature. From this reason, the measurement was performed at quasistationary states. Cooling water and oil had been cooled down before measurement started as much as possible using pumps and exchangers. The lowest temperature was limited by temperature of external cooling water in the laboratory (aprox. 20 C). Figure 5 (left) shows a warm-up process during measurement of one operation point (defined by RPM and throttle position). The aim was to achieve minimum difference between oil and water temperature. Warm-up speed was slow. Requirement was to keep maximum difference of 1 C during 100 cycles of indication. Example of warmup velocity is on Figure 5 (right). Temperature [ C] Oil temp. toil Water temp. two Evaluated points Evaluated points Time [min] Speed of heating [ C/s] Time [min] Oil Water Figure 5: Warm-up procedure (RPM = 3500 min -1, bmep = 660 kpa) 2.5 Results Complete engine characteristic has been measured. Figure 6 shows results of FMEP and oil pressure for different oil temperatures (left) and mechanical efficiency (right) Curves are interpolated with polynomial (4 th order). Results of FMEP are compared for motored and firing engine with full open throttle. FMEP is higher for motoring engine. These unexpected results were observed for all measured speeds. Difference is higher for lower oil temperature and is higher for lower speed. There are several reasons for those differences published in the literature [1]. Most of them have been eliminated. There are also many other possible reasons, such as inaccuracy in the engine torque measurement of ±2 Nm which causes difference of BMEP calculation of ±21 kpa. Further, precision of IMEP determination could be influenced by phenomenon called

7 temperature drift of the piezoelectric pressure sensor. Higher friction can be caused due to the colder engine parts at motoring. The Piston is oval-shaped and during warming-up it changes its shape to circular one. Also an exhaust valve has bigger clearance between stem and its bushing. It can explain lower difference between operation points at higher oil temperatures, where the engine parts are more warmed-up. Partial conclusion of this measurement is reduction of the FMEP in the range from 30% to 50% during warm-up of oil and cooling water from 20 to 90 C. Combustion Motoring poil-combustion Poly. (Combustion) poil-motoring Poly. (Motoring) Combustion Poly. (Combustion) FMEP [kpa] Oil pressure [kpa] Mechanical efficiency [%] Oil temperature [ C] Oil temperature [ C] Figure 6: FMEP and oil pressure (left) and mechanical efficiency (right) at full load 2500 min -1, bmep=1050kpa 2.6 Torque measurement reconstruction Accuracy of FMEP determination is influenced by accuracy of IMEP and BMEP. BMEP is calculated from engine torque see (3). Engine torque is calculated from reaction of cradle-type stator. This reaction has been measured by reversible weight mechanism see Figure 7 (left). Figure 7: Reversible weight measurement (left) and reconstruction to tenzometric measurement (right)

8 Position of a needle was sensed and converted to analog signal using potentiometer on the shaft of the needle. Estimation of accuracy of torque measurement including potentiometer and voltage source error was ± 2 Nm. This accuracy is too low especially for friction estimation on motored engine. Requirement to improve accuracy led to in-house performed reconstruction. Weight mechanism was replaced by two tensometric sensor HBM in series (Figure 7 right). Bigger sensor has nominal load 500 N, it means 225 Nm. Smaller sensor has nominal load 100 N, it means 45 Nm. The bigger one does not demand any protection against overload due its nominal load. Nevertheless, for future use endstops are prepared. Endstops are made of nuts and a screw rod. A Small sensor would work most time of its life in overload. It is protected by metal frame. Position of two nuts above and inside the frame sets specific load to adjust protection. Figure 8 (right) shows function of protection frame at static load. Nominal load is not overstepped in case of four time overload. Moreover, maximum operating force of the sensor is 150 % of nominal load it means 150 N. Mentioned protection trims operating area to ± 35 Nm. It is sufficient enough for a torque measurement at motored engine at all possible operating points sensor protection sensor protection Measured force [N] Calibrating force [N] Figure 8: Calibrating arm(left), static load of small tensometric sensor (right) Figure 8 (left) shows special arms for calibration procedure. The both sensors are calibrated together in many points. The whole measuring chain (Figure 9) is calibrated together. It means that an error of sensor, an amplifier with analog input and a measurement A/D card are taken into account. The calibrating weight is recalculated to corresponding torque. Then, a result of calibration is a table with dependency of the torque and voltage for each sensor. sensor HBM S2 connection 6-wire amplifier HBM MVD2510 analog signal 0-10V Figure 9: Tensometric measuring chain PC measurement A/D card NI PCI-6229 Up to now no experiments were carried out with this new set-up. Better accuracy is supposed to be obtained from smaller sensor. The reason is a lower hysteresis and a

9 lower influence of an ambient temperature, an eccentricity and a transverse force to sensors accuracy. 3. CONCLUSION The cost effective conditioning system has been realized for oil and cooling water. A complete engine characteristic has been measured with focus on mechanical efficiency and FMEP. The measurement at quasi-stationary states was performed. Each point of the characteristic was measured with oil temperature in range aprox C. The main result is that FMEP reduction during warm-up process of oil and cooling water (range C) is in range 30-50%. This result will be used to improve accuracy in simulation of engine behaviour while vehicle pass NEDC procedure. Two methods of FMEP measurement have been compared. Difference in results has not been sufficiently explained yet. Accuracy of engine torque measurement has been improved. Future work will be focused on improving TDC determination using the symmetry method published in [4]. REFERENCES [1] HEYWOOD, J. B.: Internal Combustion Engine Fundamentals, McGraw-Hill, Inc., 1998, pp , ISBN X [2] Takáts, M.: ITI-ONL software for acquisition of in-cylinder pressure pattern, Josef Božek Research Centre Code Library, CTU Prague, 2002 [3] Takáts, M.: INTEC software for evaluation of in-cylinder pressure record, Josef Božek Research Centre Code Library, CTU Prague, 2003 [4] Nilsson, Y. Eriksson, L.: Determining TDC Position Using Symmetry and Other Methods. SAE Technical Paper , ISSN , SAE International 2004 ACKNOWLEDGEMENT This work was supported by the Josef Božek Research Centre - project No.1M0568 at the Ministry of Education, Youth and Sports.