This paper presents results of experimental investi- gations on the friction losses of the crankshaft radial lip seals of a modern 4-cylinder diesel engine for pas- senger car applications. A two stage strip-test has been conducted on a motored engine test bed to ob- tain the friction torque of the radial lip seals. For the experimental investigations with the crankshaft seals removed from the engine, a special sealing ap- paratus has been designed and build. A wide range of tests have been performed covering the full speed range of the engine at lubricant temperatures of 70 C, 90C and 110C. The results show a depen- dency on crankshaft speed and engine media supply temperature but also revealed the presence of con- stant plateaus of friction torque over engine speed. 1 Introduction One of the crucial topics in engine development is to improve the engine eciency to cut fuel consumption and subsequently reduce emissions. Very tight emis- sion limits driven by the legislative and customers forces the automobile manufacturers to consider- ably lower the emissions of their vehicle eet (95 g CO2/km by 2021 (phased in from 2020)). For a diesel engine used in passenger cars this means a required VIRTUAL VEHICLE Research Center, Austria yVIRTUAL VEHICLE Research Center, Austria zVIRTUAL VEHICLE Research Center, Austria xAVL List GmbH, Austria,Institute of Internal Combustion Engines and Thermodynamics, Graz University of Technology, Austriafuel consumption of about 3.6 l/100 km [1]. Beside improvements of the thermodynamic eciency (e.g. operating- and combustion-processes) and electri ca- tion, friction reduction represents a very ecient and economic possibility to improve fuel consumption. In addition to the largest contributors of the frictional losses in internal combustion engines (piston assem- bly, journal bearings and valvetrain) even small con- tributors like the radial lip seals need to be analysed and improved to reach the tight emission limits. In automotive engines, the sealing of the crankshaft ends is generally ensured by radial lip seals which are installed at both ends of the engine crankshaft ( ywheel side and FEAD1side). The main functions of the radial lip seals are to prevent that dust, dirt and water gets into the engine and to avoid engine oil leakage. Therefore, the crankshaft seals need to ensure static sealing between the outer side of the ra- dial lip seal ring and the housing (e.g. crankcase) as well as dynamic and static sealing at the seal lip and shaft interface. These functions need to be guaran- teed even under severe operating conditions like high oilsump temperatures above 130C, high additive low-viscosity engine oils, sooted and diluted engine oil and tumble motion in radial and axial direction of the crankshaft ends due to vibrations from the variying load of the reciprocating pistons. [2] How- ever, modern radial lip seals for automotive applica- tions are designed for lowest possible radial contact force to maximise durability and eciency. The re- sulting radial force is caused by material expansion of the sealing body during the assembly and the spring 1Front End Auxiliary Drive 1 force of the tension spring (so-called worm or garter spring) if used. In the automotive industry it is nowa- days common to use radial lip seals without addi- tional tension springs. For automotive applications, a common value for the radial contact force related to the shaft circumference is about 0.1 N/mm. [3] To ensure proper sealing capability with low radial contact forces during dynamic operation, the radial lip seals are manufactured with macroscopic recircu- lating structures (twist) to promote a hydrodynamic sealing assistance of these surface structures. The oil supply of the crankshaft radial lip seal rings is typically done solely due to splashed oil inside the crankcase. Nevertheless, in the contact zone between the surface of the rotating crankshaft and the seal lip, a lubricating and sealing oil lm is formed. A lot of research has been performed in the past to under- stand and describe this sealing mechanism, which is a complex tribological interaction between radial lip seal, oil lm and shaft. [4], [5], [6], [7], [8], [9] The turning of the crankshaft creates shear stress in the lubrication gap which generates frictional losses. No- tably, only a minor number of publications discuss the friction losses caused by the radial shaft seals in modern engines. [10], [11], [12] To investigate the fric- tion losses of the crankshaft radial lip seals a series of tests have been conducted. Therefore, measurements have been done with installed and with removed ra- dial lip seals. To seal the engine crankcase during engine testing without the radial lip seals in place, a device has been designed and build which is able to prevent oil leakage from the engine without having contact with the crankshaft surface. The friction power losses of the crankshaft radial lip seals were measured over the full rotational speed range of the engine. As the temperature of the lu- bricant has a considerable in uence on the viscosity in the lubricated contacts and, consequently, on the tribological conditions [9], [12] in the seal lip-shaft interface, the engine was operated with di erent lu- bricant temperatures ranging from 70C to 110C.2 Testing The crankshaft seals put to test are the original radial shaft seals belonging to a modern in-line 4-cylinder, turbo charged, passenger car diesel engine with a nominal volume displacement of 2 litres and a nom- inal power of 135 kW; more technical details of the engine are listed in table 1. Table 1: Technical data of the in-line 4-cylinder tur- bocharged engine under test: Volume displacement 1995 cm3 Compression ratio 16.5:1 Bore 84 mm Stroke 90 mm Nominal torque 380 Nm Nominal Power 135 kW Maximum Speed 4600 rpm Cylinder diameter 91 mm Conrod length 138 mm Main bearing diameter 55 mm Main bearing width 25 mm Big-End bearing diameter 50 mm Big-End bearing width 24 mm Valve-train DOHC Timing drive chain Valve-train type roller-type cam follower Valves 4 per cylinder Figure 1 shows the ywheel side and the FEAD side of the crankshaft with mounted radial lip seals whereas the basic geometry and elastomer material is listed in table 2 respectively. The crankshaft seal ring on the FEAD side is shown in detail in gure 2. The seal ring construction on the ywheel side is identical. The crankshaft seal construction consists of a metal housing with an o-ring seal on the outer face of the seal ring. To seal the crankcase on the crankshaft surface, the seal ring consists of two lips. The seal- ing lip is preventing oil leakage from the oil side (crankcase) and the dust lip is preventing that dirt and water gets inside the crankcase. 2 Figure 1: Mounted crankshaft radial lip seals Table 2: Technical data of the crankshaft radial lip seals: Mounting position Flywheel side FEAD side Material FPM1FPM Outer Diameter 104 mm 68 mm Hight 9 mm 8 mm Inner Diameter 90 mm 55 mm Twist type left-hand right-hand 1Fluorinated rubber Figure 2: FEAD side crankshaft seal ring: Seal ring (left); Sectional view (right)2.1 Engine preparation and set up To separate the frictional losses of the radial lip seals from the other engine parts and components, all aux- iliary devices like oil- and water-pump, alternator, balancing shafts, vacuum pump, etc. have been re- moved from the engine or are deactived. Also, con- rods and pistons as well as the valve train timing drive (upper timing drive) have been removed. At the crankshaft, balancing weights have been installed on the crank pins to compensate the missing rotational mass forces of the conrods and to prevent excessive oil leakage from the oil supply bores. The lower tim- ing drive was mounted during the experimental in- vestigations but the high pressure fuel pump which is normally driven by the lower timing drive was de- activated. To deactivate the high pressure fuel pump, a ball bearing was mounted between the fuel pump drive shaft and chain sprocket. To prevent oil leakage from the engine crankcase dur- ing the measurements without mounted crankshaft seals, a convenient sealing at both ends of the engine is necessary. On the FEAD side, the sealing of the crankcase can be realised by using a simple plug be- cause the crankshaft shoulder is not touching the seal lip when unmounting the belt pulley (see gure 3). Figure 3: Sealing plug to prevent oil leakage on the FEAD side of the engine Proper sealing of the engine on the ywheel side without mounted crankshaft seal demands a special apparatus to prevent oil leakage from the resulting sealing gap (see gure 4). For this purpose, a custom sealing case has been designed and realised. The sealing case consists of 3 Figure 4: Flywheel side of the engine without in- stalled crankshaft seal a base plate and an oil catching chamber. The base plate is mounted on the engine crankcase and oil pan using the gearbox ange xation and is sealed with standard automotive ange sealant. The oil catch- ing chamber captures the oil which is ying o the rotating crankshaft and drive shaft ange when op- erating the engine. In addition, a bore was realised to allow the collected oil to ow back to the oil sump. The oil return bore prevents a lling of the oil catch- ing chamber and undesired e ects due to oil churn- ing of the rotating shaft during the tests. The oil catching chamber is mounted and aligned on the base plate and also sealed with standard automotive ange sealant on the front end. The catching chamber is precisely manufactured on the inner bore to prevent touching the crankshaft or drive shaft ange and to avoid oil leakage. Therefore, the bore clearance is within 0.1 mm to the crankshaft/drive shaft ange surface and is precisely aligned during the mounting procedure. The designed sealing system worked very well and no oil leakage was detected during the con- ducted measurements. Following gure 5 shows the realised sealing system on the ywheel side. The drive shaft ange connects the engine crankshaft to the torque measurement system and dynamometer with a test bed drive shaft. For the torque and engine speed measurement, a high accu- Figure 5: Realised crankshaft sealing system on the ywheel side racy torque measurement system (HBM T12) with built-in engine speed sensor is used. The connected engine crankshaft to the dynamometer is shown in gure 6. Figure 6: Crankshaft sealing system with connected crankshaft to the dynamometer To realise reproduceable thermal boundary con- ditions during the measurements, the engine media (engine oil and coolant) are supplied from external supply units. This enables a accurate control of the supply temperatures and pressures, which are mea- sured directly at the coolant circuit and oil gallery of the engine, respectively. The oil and coolant sup- 4 ply temperature was controlled with an accuracy of 0:1C. The oil and coolant pressure was controlled with an accuracy of 0.05 bar. A standard 5W30 grade automotive engine oil was used to lubricate the engine. The basic rheological properties of the used lubricant is given in table 3. Table 3: Basic rheological properties of the lubricant SAE2class 5W30 Density at 15C 853 kg/ m3 Dynamic viscosity at 40C 59.88 mPas Dynamic viscosity at 100C 9.98 mPas HTHTS3viscosity 3.57 mPas 2.2 Testing Procedure To determine the friction losses of the crankshaft ra- dial lip seals, a strip-measurement campaign was re- alised. Figure 7 shows the process steps during the experimental investigation. Figure 7: Measurement campaign process to deter- mine the frictional losses of the crankshaft seal rings The testing procedure consist of a two stage strip- measurement campaign. In the rst stage, the 2SAE: society of automotive engineers 3HTHS: high temperature high shear rate viscosity (de ned as the dynamic viscosity of the lubricant measured at 150C and at a shear rate of 106s1drag torque ( Mr Strip-Test 1 ) of the crankshaft main- bearings, the lower timing drive, the ball bearing of the fuel pump chain sprocket and the crankshaft seal- ings (FEAD side and ywheel side) was measured. The measurements were done within a crankshaft speed range of n=1000 rpm to n=4000 rpm using speed steps of n=500 rpm. In addition, the tests are conducted at three di erent engine media supply temperatures (70C, 90C and 110C) to investi- gate the in uence of the engine media supply tem- perature on the frictional losses. In the second stage, the drag torque ( Mr Strip-Test 2 ) of the engine setup from strip-test stage 1 without the crankshaft seals is measured at the same engine speeds and media supply temperatures. The resulting friction torque (Mr Crankshaft seals ) is calculated by subtraction of the measured drag torques of strip-test stage 1 and strip- test stage 2. Mr Crankshaft seals =Mr Strip-Test 1 Mr Strip-Test 2 (1) To minimize variations in torque measurement due to di erent lubricant supply temperatures, speci c focus has been spend on the lubricant supply tem- peratures when running the strip-tests. Following gure 8 shows the di erence of the oil supply temper- ature between strip-test stage 1 and strip-test stage 2 over the entire speed range for a lubricant supply temperature of 90C. The temperature variations between the strip-test stages could be maintained within 0.1C. The mea- surement uncertainty at n=1000 rpm and an engine media supply temperature of 90C isu=0.07 Nm which includes the sensitivity tolerance, linearity de- viation including hysteresis, temperature in uence, relative standard deviation of the repeatability and parasitic loads (axial force, lateral force and bending moment). At n=4000 rpm and an engine media sup- ply temperature of 90C the calculated measurement uncertainty is u=0.09 Nm. The very stable testing boundary conditions in combination with the highly accurate torque and speed measurement system are enabling experimental investigations with high repro- ducibility of the results. 5 Figure 8: Variation of oil supply temperatures 3 Results and Discussion In this section, the obtained friction losses of the crankshaft seals from the used strip-measurement procedure are discussed. Therefore, the in uence of the engine speed and the lubricant supply tem- perature on the friction losses is investigated. The strip-test measurements are carried out for 21 di er- ent measurement points for every strip-test stage (42 measurement points in total) and the resulting fric- tion torque of the crankshaft seals using equation 1 is shown in gure 9. Several interesting points can be observed from g- ure 9: The friction torque shows a dependency on the engine speed for lower engine speeds and enters a stable plateau at speci c values. At an engine me- dia supply temperature of 70C, the friction torque increases from 0.52 Nm at n=1000 rpm to 0.61 Nm at n=2000 rpm which is an increase of 0.09 Nm (17 %). For engine speeds above n=2000 rpm the friction torque reaches a constant plateau and remains con- stant at about 0.6 Nm. Further, it is interesting to note that for a lubricant supply temperature of 90C the friction torque increases from 0.42 Nm at n=1000 rpm to 0.56 Nm at n=3000 rpm which accounts to an Figure 9: Resulting friction torque of the crankshaft seals ( ywheel side and FEAD side) using equation 1 increase of 0.14 Nm or 33 %. For a lubricant supply temperature of 110C the friction torque increases from 0.35 Nm at n=1000 rpm to 0.54 Nm at n=3000 rpm which accounts to an increase of 0.19 Nm (54 %). It is also found that at higher lubricant supply tem- peratures of 90C and 110C the speci c value of engine speed - where the stable friction torque plateau begins - is shifted to higher engine speeds above n=3000 rpm. When analysing the friction torque for di erent lubri- cant supply temperatures, the lowest friction torque occurs at the highest lubricant supply temperature of 110C over the entire speed range. The highest fric- tion torque, on the other hand, occurs at the lowest lubricant supply temperature of 70C. It is interest- ing to note that the largest di erences occur at low engine speeds. At an engine speed of n=1000 rpm, the friction torque reduces from 0.52 Nm at an oil supply temperature of 70C to 0.35 Nm at an oil supply temperature of 110C which represents a re- duction of 0.17 Nm (33 %). Furthermore, the friction reduction due to higher engine media supply temper- atures is not following a linear trend. Despite that the 6 temperature di erence between the measurements at di erent engine media supply temperatures is always T=20C, the reduction of the friction torque be- tween 70C and 90C is 0.1 Nm and between 90C and 110C it is 0.07 Nm at n=1000 rpm. In addition, the advantages in the frictional losses at 110C supply temperatures is decreasing with incras- ing engine speeds. Whereas the di erence between the friction torque at 90C and 110C at an en- gine speed of n=1000 rpm is 0.07 Nm, the di erence decreases to 0.01 Nm at n=3500 rpm. Also, the dif- ference between the friction torque at 70C and 90 C at an engine speed of n=1000 rpm is 0.1 Nm, and decreases with increasing engine speed to 0.04 Nm at n=3000 rpm. A potential explanation for the frictional behavior of the radial lip seal is found in [12] who identi ed a sim- ilar trend. With increasing temperature, the radial forces of the sealing lip decreases because of changes in the rubber material properties. Additionally, the amount of oil in the lubricated contact impacts the lubrication gap and consequently the friction torque. The oil amount is dependent on crankshaft speed, due to recirculation e ects in the lubricated contact of the sealing lip. Therefore, at low engine speeds, the reduced radial force in combination with the reduced lubricant viscosity at higher temperatures explains the varying friction torque at di erent temperatures. At high engine speeds, the uniform results may be caused by the reduced amount of oil in the lubricated contact. In addition, an increase of temperature with increasing crankshaft speed can be expected which results in a further reduction of the radial force of the sealing lip. Finally, the resulting friction power losses of the crankshaft radial lip seals at di erent temperatures are analysed. In gure 10 the resulting friction power losses are plotted over engine speed for di erent engine media supply temperatures. As expected, the friction power losses show a strong increase over the crankshaft speed and is rang- ing from 37 W to 250 W depending on engine speed and lubricant supply temperature. The results show, that signi cant energy saving is possible beside large contributors like the crankshaft journal bearings [13] also at apparently small contributors of frictional losses. With an assumption of about 10 di erent Figure 10: Resulting friction power loss of the crankshaft seals ( ywheel side and FEAD side) lip seals mounted in a conventional passenger car to seal rotating shafts, the amount of frictional losses of these machine elements is signi cant [14]. For com- parison the friction power losses of the crankshaft journal bearings of a passenger car engine in motored conditions at an engine speed of n=4000 rpm and an oil supply temperature of 90C are in the magnitude of about 2 kW. 4 Conclusions The friction losses of the crankshaft seals of a modern 4- cylinder diesel engine for passenger car applica- tions have been investigated with a two stage strip- test procedure. The measurements contained tests with mounted and unmounted crankshaft seals. For the tests with unmounted crankshaft seals, a proper sealing system at both ends of the engine was de- signed and realised to prevent oil leakage during en- gine operation. A series of tests for engine speeds ranging from n=1000 rpm to n=4000 rpm and en- gine media supply temperatures of 70C, 90C and 110C have been conducted using a standard 5W30 grade automotive lubricant. 7 It was found that the highest crankshaft seal friction losses occur at the lowest lubricant supply tempera- ture. In contrast, the lowest frictional losses appear at the highest lubricant supply temperature over the entire engine speed range. This is caused by a de- creasing radial force of the crankshaft seal rings with increasing temperature in combination with the re- duced lubricant viscosity at higher temperatures. Furthermore, the results showed that the friction torque enters a stable plateau at speci c engine speeds. The transitional crankshaft speed shifts from 2000 rpm at a lubricant supply temperature of 70C to an engine speed of n=3000 rpm at 90C and 110 C. A potential explanation for this behaviour can be the reduced amount of oil in the lubricated con- tact due to the recirculation e ects of the sealing lip resulting in lower friction losses. This in combina- tion with a speed dependent temperature increase, a further decrease of the radial force of the sealing lip reduces the friction torque resulting in stable plateaus of friction torque. Finally, it was found that the friction power losses of the crankshaft seals are ranging from 37 W to 250 W depending on engine speed and lubricant supply temperature. When considering that about 10 lip shaft seals are mounted in a conventional passenger car, improvements to minimize the friction losses can make a signi cant contribution even at apparently small contributors to meet future emission limits. 5 Acknowledgement The publication was written at VIRTUAL VEHI- CLE Research Center in Graz and partially funded by the COMET K2 { Competence Centers for Excellent Technologies Programme of the Federal Ministry for Transport, Innovation and Technology (bmvit), the Federal Ministry for Digital, Business and Enterprise (bmdw), the Austrian Research Promotion Agency (FFG), the Province of Styria and the Styrian Busi- ness Promotion Agency (SFG). Furthermore, we acknowledge the partial nan- cial support of the Austrian Science Fund (FWF): P27806-N30.