Aus der Praxis für die Praxis 1 Requirements of Bearing Materials Plain bearings made of layers of different materials are widely used in combustion engines because: they require small dimensions; provide good damping and hence NVH behaviour; are easy to assemble as a split design with upper and lower half shells; and last but not least they offer economic advantages compared to other bearing concepts, such as roller bearings. Bearing products include radial sliding bearings consisting of two half shells each covering 180° of the circumference, bushings which are 360° round components, and thrust washers which are flat and used to transfer the axial loads of a shaft into the housing. Depending on their location inside the engine and their particular use, the bearings have to fulfil their purpose under quite different boundary conditions (Figure 1). Beside the technical demands, the materials also have to comply with legislative regulations. All types of bearings must be able to withstand a certain level of contamination by foreign particles without loss of function. Sources of contamination include: residuals from engine component production and assembly; particles passing through the air and oil filters during engine operation; debris from wear processes of engine parts inside the oil circuit; and ingress of debris when operating surfaces are exposed during repair and maintenance. Contaminating media can include sand, glass, dust, metallic chips and fibres. The bearings are part of a system with the shaft as their mechanical counterpart. Its surface roughness and shape are parameters which influence the performance of the bearings significantly. For this reason, the bearing material must have the potential for conditioning of the journal, to mitigate aggressive surface characteristics. 1.1 Conrod Small End Bushings The small end bushing is mechanically the highest loaded bearing component in an engine. Close to the piston, it Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 43 * Dipl.-Ing. Christian Wolf Dr.-Ing. Thierry Garnier Dipl.-Ing. Maik Wilhelm Federal-Mogul Wiesbaden GmbH, Wiesbaden Tribology and Technology of Bearings for Crankshaft-Conrod-Systems C. Wolf, T. Garnier, M. Wilhelm* Tribology and Technology of Bearings for Crankshaft-Conrod-Systems Christian Wolf, Thierry Garnier, Maik Wilhelm Federal-Mogul Wiesbaden GmbH, Wiesbaden, Germany Summary The development of combustion engines is focused particularly on the reduction of CO2 emissions. To develop more efficient engines, downsizing is used leading to increased combustion peak cylinder pressures which impact on the tribological conditions of the base engine's bearing components. Higher operating temperatures require bearing materials with improved corrosion resistance especially for the connecting rod small end bushing. Increased mechanical loads require the use of materials with sufficient fatigue resistance. This is combined with the demand for improved wear resistance to enable start-stop functionality. Beside the impact of efficiency improvements driven by other areas of the engine, the bearings themselves can also contribute to friction reduction. This is achieved directly on the one hand by optimization of hydrodynamic friction power loss and indirectly on the other hand, by minimizing the power loss caused by excessive oil drainage from the bearings increasing the amount of oil flow required from the oil pump. 1. Requirements of Bearing Materials Plain bearings made of layers of different materials are widely used in combustion engines because: they require small dimensions; provide good damping and hence NVH behaviour; are easy to assemble as a split design with upper and lower half shells; and last but not least they offer economic advantages compared to other bearing concepts, such as roller bearings. Bearing products include radial sliding bearings consisting of two half shells each covering 180° of the circumference, bushings which are 360° round components, and thrust washers which are flat and used to transfer the axial loads of a shaft into the housing. Depending on their location inside the engine and their particular use, the bearings have to fulfil their purpose under quite different boundary conditions (Fig. 1). Beside the technical demands, the materials also have to comply with legislative regulations. All types of bearings must be able to withstand a certain level of contamination by foreign particles without loss of function. Sources of contamination include: residuals from engine component production and assembly; particles passing through the air and oil filters during engine operation; debris from wear processes of engine parts inside the oil circuit; and ingress of debris when operating surfaces are exposed during repair and maintenance. Contaminating media can include sand, glass, dust, metallic chips and fibres. 1.1 Conrod Small End Bushings The small end bushing is mechanically the highest loaded bearing component in an engine. Close to the piston, it is exposed to high temperatures and often lubricated by oil used for piston cooling. The reciprocating motion of the conrod and complex rotating behaviour of the piston pin [1] lead to mostly mixed Figure 1: Spider diagram of required bearing material characteristics for different types of applications The development of combustion engines is focused particularly on the reduction of CO 2 emissions. To develop more efficient engines, downsizing is used leading to increased combustion peak cylinder pressures which impact on the tribological conditions of the base engine’s bearing components. Higher operating temperatures require bearing materials with improved corrosion resistance especially for the connecting rod small end bushing. Increased mechanical loads require the use of materials with sufficient fatigue resistance. This is combined with the demand for improved wear resistance to enable start-stop functionality. Beside the impact of efficiency improvements driven by other areas of the engine, the bearings themselves can also contribute to friction reduction. This is achieved directly on the one hand by optimization of hydrodynamic friction power loss and indirectly on the other hand, by minimizing the power loss caused by excessive oil drainage from the bearings increasing the amount of oil flow required from the oil pump. Keywords Material, Bearing Design, Friction, Oil Flow, Hydrodynamics, Viscosity, Power Loss Optimization Abstract T+S_6_16 17.10.16 17: 01 Seite 43 Aus der Praxis für die Praxis is exposed to high temperatures and often lubricated by oil used for piston cooling. The reciprocating motion of the conrod and complex rotating behaviour of the piston pin [1] lead to mostly mixed friction conditions, so high wear resistance of the material is important. Temperature and high local pressure contribute to a corrosive environment in combination with the properties of the lubricating oil. Higher temperature also means locally lower oil viscosity, calling for high seizure resistance at the same time. Today, most automotive and even heavy duty truck engines lubricate the small end by oil splash. Cu-based materials are used for small end bushings, as they deliver the high fatigue strength needed. The materials can be solid or bi-metal, consisting of bearing alloy sintered or cast on a steel backing (Figure 2). 1.2 Conrod Big End Bearings Following the flow of combustion forces through the crank train, the load of the conrod is transferred to the crankshaft by the conrod bearings. They work under hydrodynamic conditions with unidirectional sliding speed. Lubrication and cooling takes place by filtered oil from the pressurized oil circuit. The projected surface of the rod bearing is larger than the one of the small end bushing, so the mechanical load is lower; however the upper rod bearing is the highest loaded bearing shell in the engine. Additionally, it is exposed to high sliding speed. Due to the high loads, the lubricating oil film can be very thin intermittently, well below half a micron, as hydrodynamic calculations show. Thus, wear resistance is an important requirement for the upper rod bearings. Good sliding properties are required to overcome local short term interruption of the oil film, in order to control the risk of overheating and consequently seizure (Figure 3). In modern Diesel engines for automotive and heavy duty applications, bronze based trimetal compound materials are used for the upper conrod shell. The cap side bearing is much less highly loaded and usually can be equipped with a softer bi-metallic material, providing high embedability for foreign particles; by this combination, the delicate upper bearings can be protected from contamination. In gasoline engines, the specific load in the upper conrod bearing is lower compared to Diesel engines; so high strength Al-alloys in form of steel backed bi-metals can be used. Their application can be extended by polymer coating, such as IROX ® . The development of gasoline engines with turbo charging and direct injection is leading to higher combustion peak pressures. In the future, this may require more frequent use of higher load capacity materials based on Copper substrates for upper rod bearings in gasoline engines. At high engine speed, the operating forces lead to dynamic deformations of the conrod big end housing bore. The bearings must be able to adapt to the actual shape of the housing to prevent the formation of any gaps in the assembly. 1.3 Crankshaft Main Bearings The combustion force transferred from the conrod bearings is typically distributed to two main bearings. These 44 Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 Figure 2: Cu-based bi-metal materials for small end bushings friction conditions, so high wear resistance of the material is important. Temperature and high local pressure contribute to a corrosive environment in combination with the properties of the lubricating oil. Higher temperature also means locally lower oil viscosity, calling for high seizure resistance at the same time. Today, most automotive and even heavy duty truck engines lubricate the small end by oil splash. Cu-based materials are used for small end bushings, as they deliver the high fatigue strength needed. The materials can be solid or bi-metal, consisting of bearing alloy sintered or cast on a steel backing (Fig. 1). LF-5, CuSn10Bi3 sintered LF-8, CuSn8Ni sintered G-149, CuAl8 cast RC-9, CuSn6Ni9 solid Fig. 1: Cu-based bi-metal materials for small end bushings 1.2 Conrod Big End Bearings Following the flow of combustion forces through the crank train, the load of the conrod is transferred to the crankshaft by the conrod bearings. They work under hydrodynamic conditions with unidirectional sliding speed. Lubrication and cooling takes place by filtered oil from the pressurized oil circuit. The projected surface of the rod bearing is larger than the one of the small end bushing, so the mechanical load is lower; however the upper rod bearing is the highest loaded bearing shell in the engine. Additionally, it is exposed to high sliding speed. Due to the high loads, the lubricating oil film can be very thin intermittently, well below half a micron, as hydrodynamic calculations show. Thus, wear resistance is an important requirement for the upper rod bearings. Good sliding properties are required to overcome local short term interruption of the oil film, in order to control the risk of overheating and consequently seizure (Fig.2). In modern Diesel engines for automotive and heavy duty applications, bronze based tri-metal compound materials are used for the upper conrod shell. The cap side bearing is much less highly loaded and usually can be equipped with a softer bi-metallic material, providing high embed-ability for foreign particles; by this combination, the delicate upper bearings can be protected from contamination. In gasoline engines, the specific load in the upper conrod bearing is lower compared to Diesel engines; so high strength Al-alloys in form of steel backed bimetals can be used. Their application can be extended by polymer coating, such as IROX ® . The development of gasoline engines with turbo charging and direct injection is leading to higher combustion peak pressures. In the future, this may require more frequent use of higher load capacity materials based on Copper substrates for upper rod bearings in gasoline engines. At high engine speed, the operating forces lead to dynamic deformations of the conrod big end housing bore. The bearings must be able to adapt to the actual shape of the housing to prevent the formation of any gaps in the assembly. Fig. 2: Micro-sections of a selection of high strength materials suitable for upper conrod big end bearing applications 1.3 Crankshaft Main Bearings The combustion force transferred from the conrod bearings is typically distributed to two main bearings. These often have comparable or even larger dimensions than the conrod bearings, so the demand for load capacity in this position is reduced. Being closer to the source of pressurized oil, the oil temperature of the main bearings is lower than in the rod bearings; part of the friction power loss in the main bearings increases the temperature of the oil as it passes through the upper main bearing groove and via the crankshaft into the rod bearing. In general, Al-alloys are suitable to carry the mechanical load. However, more complicated lubrication situafriction conditions, so high wear resistance of the material is important. Temperature and high local pressure contribute to a corrosive environment in combination with the properties of the lubricating oil. Higher temperature also means locally lower oil viscosity, calling for high seizure resistance at the same time. Today, most automotive and even heavy duty truck engines lubricate the small end by oil splash. Cu-based materials are used for small end bushings, as they deliver the high fatigue strength needed. The materials can be solid or bi-metal, consisting of bearing alloy sintered or cast on a steel backing (Fig. 1). LF-5, CuSn10Bi3 sintered LF-8, CuSn8Ni sintered G-149, CuAl8 cast RC-9, CuSn6Ni9 solid Fig. 1: Cu-based bi-metal materials for small end bushings 1.2 Conrod Big End Bearings Following the flow of combustion forces through the crank train, the load of the conrod is transferred to the crankshaft by the conrod bearings. They work under hydrodynamic conditions with unidirectional sliding speed. Lubrication and cooling takes place by filtered oil from the pressurized oil circuit. The projected surface of the rod bearing is larger than the one of the small end bushing, so the mechanical load is lower; however the upper rod bearing is the highest loaded bearing shell in the engine. Additionally, it is exposed to high sliding speed. Due to the high loads, the lubricating oil film can be very thin intermittently, well below half a micron, as hydrodynamic calculations show. Thus, wear resistance is an important requirement for the upper rod bearings. Good sliding properties are required to overcome local short term interruption of the oil film, in order to control the risk of overheating and consequently seizure (Fig.2). In modern Diesel engines for automotive and heavy duty applications, bronze based tri-metal compound materials are used for the upper conrod shell. The cap side bearing is much less highly loaded and usually can be equipped with a softer bi-metallic material, providing high embed-ability for foreign particles; by this combination, the delicate upper bearings can be protected from contamination. In gasoline engines, the specific load in the upper conrod bearing is lower compared to Diesel engines; so high strength Al-alloys in form of steel backed bimetals can be used. Their application can be extended by polymer coating, such as IROX ® . The development of gasoline engines with turbo charging and direct injection is leading to higher combustion peak pressures. In the future, this may require more frequent use of higher load capacity materials based on Copper substrates for upper rod bearings in gasoline engines. At high engine speed, the operating forces lead to dynamic deformations of the conrod big end housing bore. The bearings must be able to adapt to the actual shape of the housing to prevent the formation of any gaps in the assembly. Fig. 2: Micro-sections of a selection of high strength materials suitable for upper conrod big end bearing applications 1.3 Crankshaft Main Bearings The combustion force transferred from the conrod bearings is typically distributed to two main bearings. These often have comparable or even larger dimensions than the conrod bearings, so the demand for load capacity in this position is reduced. Being closer to the source of pressurized oil, the oil temperature of the main bearings is lower than in the rod bearings; part of the friction power loss in the main bearings increases the temperature of the oil as it passes through the upper main bearing groove and via the crankshaft into the rod bearing. In general, Al-alloys are suitable to carry the mechanical load. However, more complicated lubrication situa- Figure 3: Micro-sections of a selection of high strength materials suitable for upper conrod big end bearing applications friction conditions, so high wear resistance of the material is important. Temperature and high local pressure contribute to a corrosive environment in combination with the properties of the lubricating oil. Higher temperature also means locally lower oil viscosity, calling for high seizure resistance at the same time. Today, most automotive and even heavy duty truck engines lubricate the small end by oil splash. Cu-based materials are used for small end bushings, as they deliver the high fatigue strength needed. The materials can be solid or bi-metal, consisting of bearing alloy sintered or cast on a steel backing (Fig. 1). LF-5, CuSn10Bi3 LF-8, CuSn8Ni G-149, CuAl8 RC-9, CuSn6Ni9 Fig. 1: Cu-based bi-metal materials for small end bushings 1.2 Conrod Big End Bearings Following the flow of combustion forces through the crank train, the load of the conrod is transferred to the crankshaft by the conrod bearings. They work under hydrodynamic conditions with unidirectional sliding speed. Lubrication and cooling takes place by filtered oil from the pressurized oil circuit. The projected surface of the rod bearing is larger than the one of the small end bushing, so the mechanical load is lower; however the upper rod bearing is the highest loaded bearing shell in the engine. Additionally, it is exposed to high sliding speed. Due to the high loads, the lubricating oil film can be very thin intermittently, well below half a micron, as hydrodynamic calculations show. Thus, wear resistance is an important requirement for the upper rod bearings. Good sliding properties are required to overcome local short term interruption of the oil film, in order to control the risk of overheating and consequently seizure (Fig.2). In modern Diesel engines for automotive and heavy duty applications, bronze based tri-metal compound materials are used for the upper conrod shell. The cap side bearing is much less highly loaded and usually can be equipped with a softer bi-metallic material, providing high embed-ability for foreign particles; by this combination, the delicate upper bearings can be protected from contamination. In gasoline engines, the specific load in the upper conrod bearing is lower compared to Diesel engines; so high strength Al-alloys in form of steel backed bimetals can be used. Their application can be extended by polymer coating, such as IROX ® . The development of gasoline engines with turbo charging and direct injection is leading to higher combustion peak pressures. In the future, this may require more frequent use of higher load capacity materials based on Copper substrates for upper rod bearings in gasoline engines. At high engine speed, the operating forces lead to dynamic deformations of the conrod big end housing bore. The bearings must be able to adapt to the actual shape of the housing to prevent the formation of any gaps in the assembly. F-211 St/ AlSnNi/ IROX ® G-611 St/ CuSnBi/ IROX ® A-280, St/ AlSnNi G-488 St/ CuNiSi/ Ni/ NiSn/ SnCu electro plating Fig. 2: Micro-sections of a selection of high strength materials suitable for upper conrod big end bearing applications 1.3 Crankshaft Main Bearings The combustion force transferred from the conrod bearings is typically distributed to two main bearings. These often have comparable or even larger dimensions than the conrod bearings, so the demand for load capacity in this position is reduced. Being closer to the source of pressurized oil, the oil temperature of the main bearings is lower than in the rod bearings; part of the friction power loss in the main bearings increases the temperature of the oil as it passes through the upper main bearing groove and via the crankshaft into the rod bearing. In general, Al-alloys are suitable to carry the mechanical load. However, more complicated lubrication situafriction conditions, so high wear resistance of the material is important. Temperature and high local pressure contribute to a corrosive environment in combination with the properties of the lubricating oil. Higher temperature also means locally lower oil viscosity, calling for high seizure resistance at the same time. Today, most automotive and even heavy duty truck engines lubricate the small end by oil splash. Cu-based materials are used for small end bushings, as they deliver the high fatigue strength needed. The materials can be solid or bi-metal, consisting of bearing alloy sintered or cast on a steel backing (Fig. 1). LF-5, CuSn10Bi3 LF-8, CuSn8Ni G-149, CuAl8 RC-9, CuSn6Ni9 Fig. 1: Cu-based bi-metal materials for small end bushings 1.2 Conrod Big End Bearings Following the flow of combustion forces through the crank train, the load of the conrod is transferred to the crankshaft by the conrod bearings. They work under hydrodynamic conditions with unidirectional sliding speed. Lubrication and cooling takes place by filtered oil from the pressurized oil circuit. The projected surface of the rod bearing is larger than the one of the small end bushing, so the mechanical load is lower; however the upper rod bearing is the highest loaded bearing shell in the engine. Additionally, it is exposed to high sliding speed. Due to the high loads, the lubricating oil film can be very thin intermittently, well below half a micron, as hydrodynamic calculations show. Thus, wear resistance is an important requirement for the upper rod bearings. Good sliding properties are required to overcome local short term interruption of the oil film, in order to control the risk of overheating and consequently seizure (Fig.2). In modern Diesel engines for automotive and heavy duty applications, bronze based tri-metal compound materials are used for the upper conrod shell. The cap side bearing is much less highly loaded and usually can be equipped with a softer bi-metallic material, providing high embed-ability for foreign particles; by this combination, the delicate upper bearings can be protected from contamination. In gasoline engines, the specific load in the upper conrod bearing is lower compared to Diesel engines; so high strength Al-alloys in form of steel backed bimetals can be used. Their application can be extended by polymer coating, such as IROX ® . The development of gasoline engines with turbo charging and direct injection is leading to higher combustion peak pressures. In the future, this may require more frequent use of higher load capacity materials based on Copper substrates for upper rod bearings in gasoline engines. At high engine speed, the operating forces lead to dynamic deformations of the conrod big end housing bore. The bearings must be able to adapt to the actual shape of the housing to prevent the formation of any gaps in the assembly. F-211 St/ AlSnNi/ IROX ® G-611 St/ CuSnBi/ IROX ® Fig. 2: Micro-sections of a selection of high strength materials suitable for upper conrod big end bearing applications 1.3 Crankshaft Main Bearings The combustion force transferred from the conrod bearings is typically distributed to two main bearings. These often have comparable or even larger dimensions than the conrod bearings, so the demand for load capacity in this position is reduced. Being closer to the source of pressurized oil, the oil temperature of the main bearings is lower than in the rod bearings; part of the friction power loss in the main bearings increases the temperature of the oil as it passes through the upper main bearing groove and via the crankshaft into the rod bearing. In general, Al-alloys are suitable to carry the mechanical load. However, more complicated lubrication situafriction conditions, so high wear resistance of the material is important. Temperature and high local pressure contribute to a corrosive environment in combination with the properties of the lubricating oil. Higher temperature also means locally lower oil viscosity, calling for high seizure resistance at the same time. Today, most automotive and even heavy duty truck engines lubricate the small end by oil splash. Cu-based materials are used for small end bushings, as they deliver the high fatigue strength needed. The materials can be solid or bi-metal, consisting of bearing alloy sintered or cast on a steel backing (Fig. 1). LF-5, CuSn10Bi3 LF-8, CuSn8Ni G-149, CuAl8 RC-9, CuSn6Ni9 Fig. 1: Cu-based bi-metal materials for small end bushings 1.2 Conrod Big End Bearings Following the flow of combustion forces through the crank train, the load of the conrod is transferred to the crankshaft by the conrod bearings. They work under hydrodynamic conditions with unidirectional sliding speed. Lubrication and cooling takes place by filtered oil from the pressurized oil circuit. The projected surface of the rod bearing is larger than the one of the small end bushing, so the mechanical load is lower; however the upper rod bearing is the highest loaded bearing shell in the engine. Additionally, it is exposed to high sliding speed. Due to the high loads, the lubricating oil film can be very thin intermittently, well below half a micron, as hydrodynamic calculations show. Thus, wear resistance is an important requirement for the upper rod bearings. Good sliding properties are required to overcome local short term interruption of the oil film, in order to control the risk of overheating and consequently seizure (Fig.2). In modern Diesel engines for automotive and heavy duty applications, bronze based tri-metal compound materials are used for the upper conrod shell. The cap side bearing is much less highly loaded and usually can be equipped with a softer bi-metallic material, providing high embed-ability for foreign particles; by this combination, the delicate upper bearings can be protected from contamination. In gasoline engines, the specific load in the upper conrod bearing is lower compared to Diesel engines; so high strength Al-alloys in form of steel backed bimetals can be used. Their application can be extended by polymer coating, such as IROX ® . The development of gasoline engines with turbo charging and direct injection is leading to higher combustion peak pressures. In the future, this may require more frequent use of higher load capacity materials based on Copper substrates for upper rod bearings in gasoline engines. At high engine speed, the operating forces lead to dynamic deformations of the conrod big end housing bore. The bearings must be able to adapt to the actual shape of the housing to prevent the formation of any gaps in the assembly. G-444 St/ CuNiSi/ Ni/ CuSn electro plating G-499 St/ CuNiSi/ NiCr/ AlSn Sputter; PVD plating Fig. 2: Micro-sections of a selection of high strength materials suitable for upper conrod big end bearing applications 1.3 Crankshaft Main Bearings The combustion force transferred from the conrod bearings is typically distributed to two main bearings. These often have comparable or even larger dimensions than the conrod bearings, so the demand for load capacity in this position is reduced. Being closer to the source of pressurized oil, the oil temperature of the main bearings is lower than in the rod bearings; part of the friction power loss in the main bearings increases the temperature of the oil as it passes through the upper main bearing groove and via the crankshaft into the rod bearing. In general, Al-alloys are suitable to carry the mechanical load. However, more complicated lubrication situa- T+S_6_16 17.10.16 17: 01 Seite 44 Aus der Praxis für die Praxis often have comparable or even larger dimensions than the conrod bearings, so the demand for load capacity in this position is reduced. Being closer to the source of pressurized oil, the oil temperature of the main bearings is lower than in the rod bearings; part of the friction power loss in the main bearings increases the temperature of the oil as it passes through the upper main bearing groove and via the crankshaft into the rod bearing. In general, Al-alloys are suitable to carry the mechanical load. However, more complicated lubrication situations can occur, for example due to start stop functionality. This may introduce the need for additional wear resistance by use of Si-containing Al-alloys or even polymer coatings which have outstandingly good wear properties. Compared to rod bearings, the main bearings are more strongly connected to each other by the crankshaft and the common housing of the engine block. Form deviations from production tolerances, static deformations due to thermal effects and dynamic deformations under operating loads can lead to additional local stress. To ensure a durable design, the bearing surfaces need to provide conformability under these conditions, which they achieve by slightly changing their shape elastically according to the operating conditions. Stress peaks can also be limited or eliminated by the wear during running in. In this respect, Al-alloys are very suitable (Figure 4). However for long term durability, Cu-based materials are also in use especially in heavy duty applications, and in some automotive main bearings. They are the same as those specified for the upper conrod bearing (Figure 3). The sliding speed in the main bearings is comparable to the rod bearings. 1.4 Crankshaft Axial Bearings The axial bearings keep the crankshaft in position and prevent the radial bearings from making contact with the journal fillet radii at their axial side faces. During engine operation, vibration and clutch forces act axially against the crankshaft, and are balanced by the thrust washers and transferred into the engine block. Beside individual washers, in some cases washers and radial bearings are connected to form flanged bearings. Modern concepts use three piece systems allowing the application of different materials for the radial and the axial parts. Connection is achieved by mechanical clamping or laser welding with and without a predetermined breaking point. Conventional single piece flange bearings are becoming less important. Compared to the radial bearings, the loads in the axial bearings are very small; they are caused by gear shift forces and guiding effects inside the engine. However, breakage of thrust washers can occur due to imperfections in the support behind them, such as steps in the housing or debris between housing and steel backing. The axial lubrication gap is defined by the two almost parallel thrust faces of crankshaft and washer. From the hydrodynamic point of view, the lubrication gap geometry can be improved by the introduction of a ramp and flat design, providing additional convergent areas. The axial bearings are lubricated by the oil draining from their associated radial bearing, entering at the inner diameter and through grooves across the sliding surfaces. The grooves also prevent a local restriction of the oil flow from the radial bearings, to avoid overheating. Dedicated materials for the thrust washers are Al-alloys with high tin content to provide good self lubrication and sliding properties. To limit wear in start stop applications, Silicon in the Al-alloy is an important wear performance enhancer. If such an AlSnSi alloy is insufficient, the washers also can be plated with a wear protecting polymer coating (Figure 4). 2 Bearings Contribution to Engine Friction Looking at the flow of energy in an engine, a considerable part of the mechanical work is converted into friction (Figure 5). The crankshaft bearings contribute significantly to this loss. The level of friction power loss depends on the operating conditions and the engine type. Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 45 Figure 4: Some high performance materials especially suitable for main bearing applications in automotive Diesel and gasoline engines tions can occur, for example due to start stop functionality. This may introduce the need for additional wear resistance by use of Si-containing Al-alloys or even polymer coatings which have outstandingly good wear properties. Compared to rod bearings, the main bearings are more strongly connected to each other by the crankshaft and the common housing of the engine block. Form deviations from production tolerances, static deformations due to thermal effects and dynamic deformations under operating loads can lead to additional local stress. To ensure a durable design, the bearing surfaces need to provide conformability under these conditions, which they achieve by slightly changing their shape elastically according to the operating conditions. Stress peaks can also be limited or eliminated by the wear during running in. In this respect, Al-alloys are very suitable (Fig. 3). However for long term durability, Cu-based materials are also in use especially in heavy duty applications, and in some automotive main bearings. They are the same as those specified for the upper conrod bearing (Fig.2). The sliding speed in the main bearings is comparable to the rod bearings. A-590, St/ AlSnSi F - 411, St/ AlSnSiCu/ IROX ® A-370, St/ AlSnCu A-490, St/ AlSnSiCu 1.4 Crankshaft Axial Bearings The axial bearings keep the crankshaft in position and prevent the radial bearings from making contact with the journal fillet radii at their axial side faces. During engine operation, vibration and clutch forces act axially against the crankshaft, and are balanced by the thrust washers and transferred into the engine block. Beside individual washers, in some cases washers and radial bearings are connected to form flanged bearings. Modern concepts use three piece systems allowing the application of different materials for the radial and the axial parts. Connection is achieved by mechanical clamping or laser welding with and without a predetermined breaking point. Conventional single piece flange bearings are becoming less important. Compared to the radial bearings, the loads in the axial bearings are very small; they are caused by gear shift forces and guiding effects inside the engine. However, breakage of thrust washers can occur due to imperfections in the support behind them, such as steps in the housing or debris between housing and steel backing. The axial lubrication gap is defined by the two almost parallel thrust faces of crankshaft and washer. From the hydrodynamic point of view, the lubrication gap geometry can be improved by the introduction of a ramp and flat design, providing additional convergent areas. The axial bearings are lubricated by the oil draining from their associated radial bearing, entering at the inner diameter and through grooves across the sliding surfaces. The grooves also prevent a local restriction of the oil flow from the radial bearings, to avoid overheating. Dedicated materials for the thrust washers are Alalloys with high Tin content to provide good self lubrication and sliding properties. To limit wear in start stop applications, Silicon in the Al-alloy is an important wear performance enhancer. If such an AlSnSi alloy is insufficient, the washers also can be plated with a wear protecting polymer coating (Fig. 3). 2. Bearings Contribution to Engine Friction Looking at the flow of energy in an engine, a considerable part of the mechanical work is converted into friction (Fig. 4). The crankshaft bearings contribute significantly to this loss. The level of friction power loss depends on the operating conditions and the engine type. Optimization efforts must take into account the relevance and frequency of occurrence of particular operating conditions. For this reason, part load conditions at intermediate engine speeds are also of interest. As an example, if the friction work of the crankshaft group is about 22% of the mechanical friction work, this represents about 2% of the energy input from the fuel. The friction power loss of bearings is driven by the direct loss caused by viscous effects inside the bearing, and extended by the indirect loss caused due to parasitic oil flow from the bearings, which needs to be compensated by the oil supply from the oil pump. tions can occur, for example due to start stop functionality. This may introduce the need for additional wear resistance by use of Si-containing Al-alloys or even polymer coatings which have outstandingly good wear properties. Compared to rod bearings, the main bearings are more strongly connected to each other by the crankshaft and the common housing of the engine block. Form deviations from production tolerances, static deformations due to thermal effects and dynamic deformations under operating loads can lead to additional local stress. To ensure a durable design, the bearing surfaces need to provide conformability under these conditions, which they achieve by slightly changing their shape elastically according to the operating conditions. Stress peaks can also be limited or eliminated by the wear during running in. In this respect, Al-alloys are very suitable (Fig. 3). However for long term durability, Cu-based materials are also in use especially in heavy duty applications, and in some automotive main bearings. They are the same as those specified for the upper conrod bearing (Fig.2). The sliding speed in the main bearings is comparable to the rod bearings. A-590, St/ AlSnSi F-411, St/ AlSnSiCu/ IROX ® 1.4 Crankshaft Axial Bearings The axial bearings keep the crankshaft in position and prevent the radial bearings from making contact with the journal fillet radii at their axial side faces. During engine operation, vibration and clutch forces act axially against the crankshaft, and are balanced by the thrust washers and transferred into the engine block. Beside individual washers, in some cases washers and radial bearings are connected to form flanged bearings. Modern concepts use three piece systems allowing the application of different materials for the radial and the axial parts. Connection is achieved by mechanical clamping or laser welding with and without a predetermined breaking point. Conventional single piece flange bearings are becoming less important. Compared to the radial bearings, the loads in the axial bearings are very small; they are caused by gear shift forces and guiding effects inside the engine. However, breakage of thrust washers can occur due to imperfections in the support behind them, such as steps in the housing or debris between housing and steel backing. The axial lubrication gap is defined by the two almost parallel thrust faces of crankshaft and washer. From the hydrodynamic point of view, the lubrication gap geometry can be improved by the introduction of a ramp and flat design, providing additional convergent areas. The axial bearings are lubricated by the oil draining from their associated radial bearing, entering at the inner diameter and through grooves across the sliding surfaces. The grooves also prevent a local restriction of the oil flow from the radial bearings, to avoid overheating. Dedicated materials for the thrust washers are Alalloys with high Tin content to provide good self lubrication and sliding properties. To limit wear in start stop applications, Silicon in the Al-alloy is an important wear performance enhancer. If such an AlSnSi alloy is insufficient, the washers also can be plated with a wear protecting polymer coating (Fig. 3). 2. Bearings Contribution to Engine Friction Looking at the flow of energy in an engine, a considerable part of the mechanical work is converted into friction (Fig. 4). The crankshaft bearings contribute significantly to this loss. The level of friction power loss depends on the operating conditions and the engine type. Optimization efforts must take into account the relevance and frequency of occurrence of particular operating conditions. For this reason, part load conditions at intermediate engine speeds are also of interest. As an example, if the friction work of the crankshaft group is about 22% of the mechanical friction work, this represents about 2% of the energy input from the fuel. The friction power loss of bearings is driven by the direct loss caused by viscous effects inside the bearing, and extended by the indirect loss caused due to parasitic oil flow from the bearings, which needs to be compensated by the oil supply from the oil pump. T+S_6_16 17.10.16 17: 01 Seite 45 Aus der Praxis für die Praxis Optimization efforts must take into account the relevance and frequency of occurrence of particular operating conditions. For this reason, part load conditions at intermediate engine speeds are also of interest. As an example, if the friction work of the crankshaft group is about 22 % of the mechanical friction work, this represents about 2 % of the energy input from the fuel. The friction power loss of bearings is driven by the direct loss caused by viscous effects inside the bearing, and extended by the indirect loss caused due to parasitic oil flow from the bearings, which needs to be compensated by the oil supply from the oil pump. 3 Direct Bearing Friction Power Loss The friction of sliding bearings depends on the actual working situation of the bearing system, which includes housing, bearings, oil and shaft journals. The condition can be described in a simplified way by the well known Stribeck curve (Figure 6). At standstill, the coefficient of friction is at its largest (Coulomb coefficient). Once the shaft rotates with low sliding speed, the system operates under boundary conditions. Here, the surface asperities interact strongly, producing abrasive wear. Bearings must not operate for prolonged periods in this Area I. Increasing the sliding speed leads to a reduction in the coefficient of friction. The oil begins to separate the surfaces, even though the system still shows frequent asperity contact in this Area II of mixed friction. Start stop functionality increases the number of these transitions through the boundary and mixed lubrication areas. Today some applications have a design target of 1 or 2 million start processes, compared to between 40 and 70 thousand lifetime starts in the past. This can become the dominant driver for wear, for example in main bearings exposed to initial static preload by belt drives or increased gravity forces. Further increase of the angular speed ω leads into Area III: the hydrodynamic effect produces enough pressure in the convergent wedge of the lubrication gap to separate the surfaces of bearing and shaft and to balance the external forces on the system. This load is described as specific load p. The friction is now driven by the shear stress of the oil, produced by the sliding speed and the dynamic oil viscosity η; therefore this situation is called fluid friction. It is the preferred area of operation for hydrodynamic bearing systems. Smaller bearing operating numbers in the fluid friction regime lead to reduced friction coefficients. Reduced sliding speed (down-speeding) and increased specific load (down-sizing) both have a positive impact on friction. 3.1 Influence of Oil Choosing oil with lower viscosity also reduces the friction coefficient. This is attractive, because the improvement also applies to other areas, such as piston friction or the pumping and flow resistance of oil through the oil galleries, and often leads to a directly measurable reduction in engine fuel consumption. 46 Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 3. Direct Bearing Friction Power Loss The friction of sliding bearings depends on the actual working situation of the bearing system, which includes housing, bearings, oil and shaft journals. The condition can be described in a simplified way by the well known Stribeck curve (Fig. 5). At standstill, the coefficient of friction is at its largest (Coulomb coefficient). Once the shaft rotates with low sliding speed, the system operates under boundary conditions. Here, the surface asperities interact strongly, producing abrasive wear. Bearings must not operate for prolonged periods in this Area I. Increasing the sliding speed leads to a reduction in the coefficient of friction. The oil begins to separate the surfaces, even though the system still shows frequent asperity contact in this Area II of mixed friction. Start stop functionality increases the number of these transitions through the boundary and mixed lubrication areas. Today some applications have a design target of 1 or 2 million start processes, compared to between 40 and 70 thousand lifetime starts in the past. This can become the dominant driver for wear, for example in main bearings exposed to initial static preload by belt drives or increased gravity forces. Further increase of the angular speed w leads into Area III: the hydrodynamic effect produces enough pressure in the convergent wedge of the lubrication gap to separate the surfaces of bearing and shaft and to balance the external forces on the system. This load is described as specific load p. The friction is now driven by the shear stress of the oil, produced by the sliding speed and the dynamic oil viscosity h; therefore this situation is called fluid friction. It is the preferred area of operation for hydrodynamic bearing systems. Smaller bearing operating numbers in the fluid friction regime lead to reduced friction coefficients. Reduced sliding speed (down-speeding) and increased specific load (down-sizing) both have a positive impact on friction. Fig. 5: Stribeck Curve. w: angular speed, h: dynamic oil viscosity, p: specific load on bearing 3.1 Influence of Oil Choosing oil with lower viscosity also reduces the friction coefficient. This is attractive, because the improvement also applies to other areas, such as piston friction or the pumping and flow resistance of oil through the oil galleries, and often leads to a directly measurable reduction in engine fuel consumption. At the same time the operating number wh/ p decreases and the operating condition of the bearing approaches or enters the mixed lubrication regime, because lower viscosity leads to a smaller oil film thickness (Fig. 5). For the bearings, very low viscosity may bring the risk of local loss of oil film and hard contact with the journal - even seizure can occur if the oil and bearing design is not thoroughly optimized. One method for fine tuning is the analysis of the lubrication film with sophisticated software tools. Often RHD (Rigid Hydro-Dynamic) models show the positive impact of lower viscosity oil on friction power loss (Fig. 6), providing acceptable minimum oil film thickness for all engine operating conditions. An advantage of the RHD calculation is the relatively limited amount of input data required. During the early stages of engine development, when more precise data from measurements are unavailable, this approach makes sense, but can be limited in its scope. Calculations with Elastic Hydro-Dynamic (EHD) models take into account the same reduction of the viscosity. Additionally, the deflections of the solid components reveal critical conditions for the same engine (Fig. 7), which would be hidden using RHD models. Effectively, the results show significant differences: At low engine speed, high combustion pressure produces asperity contact and increases the friction power loss on the rod bearings; main bearings show improvement; At intermediate and high engine speed, the deflection of the crankshaft leads to edge loading and asperity contacts which increase the friction power loss of the main bearings; rod bearings have better hydrodynamic Figure 5: Distribution of total fuel energy to single uses and losses, given for a 2,0 l Inline 4 Cylinder GDI TC gasoline engine [2] Fig. 4: Distribution of total fuel energy to single uses and losses, given for a 2,0l Inline 4 Cylinder GDI TC gasoline engine [2] 3. Direct Bearing Friction Power Loss The friction of sliding bearings depends on the actual working situation of the bearing system, which includes housing, bearings, oil and shaft journals. The condition can be described in a simplified way by the well known Stribeck curve (Fig. 5). At standstill, the coefficient of friction is at its largest (Coulomb coefficient). Once the shaft rotates with low sliding speed, the system operates under boundary conditions. Here, the surface asperities interact strongly, producing abrasive wear. Bearings must not operate for prolonged periods in this Area I. Increasing the sliding speed leads to a reduction in the coefficient of friction. The oil begins to separate the surfaces, even though the system still shows frequent asperity contact in this Area II of mixed friction. Start stop functionality increases the number of these transitions through the boundary and mixed lubrication areas. Today some applications have a design target of 1 or 2 million start processes, compared to between 40 and 70 thousand lifetime starts in the past. This can become the dominant driver for wear, for example in main bearings exposed to initial static preload by belt drives or increased gravity forces. Further increase of the angular speed w leads into Area III: the hydrodynamic effect produces enough pressure in the convergent wedge of the lubrication gap to separate the surfaces of bearing and shaft and to balance the external forces on the system. This load is described as specific load p. The friction is now driven by the shear stress of the oil, produced by the sliding speed and the dynamic oil viscosity h; therefore this situation is called fluid friction. It is the preferred area of operation for hydrodynamic bearing systems. Smaller bearing operating numbers in the fluid friction regime lead to reduced friction coefficients. Reduced sliding speed (down-speeding) and increased specific load (down-sizing) both have a positive impact on friction. Fig. 5: Stribeck Curve. w: angular speed, h: dynamic oil viscosity, p: specific load on bearing 3.1 Influence of Oil Choosing oil with lower viscosity also reduces the friction coefficient. This is attractive, because the improvement also applies to other areas, such as piston friction or the pumping and flow resistance of oil through the oil galleries, and often leads to a directly measurable reduction in engine fuel consumption. At the same time the operating number wh/ p decreases and the operating condition of the bearing approaches or enters the mixed lubrication regime, because lower viscosity leads to a smaller oil film thickness (Fig. 5). For the bearings, very low viscosity may bring the risk of local loss of oil film and hard contact with the journal - even seizure can occur if the oil and bearing design is not thoroughly optimized. One method for fine tuning is the analysis of the lubrication film with sophisticated software tools. Often RHD (Rigid Hydro-Dynamic) models show the positive impact of lower viscosity oil on friction power loss (Fig. 6), providing acceptable minimum oil film thickness for all engine operating conditions. An advantage of the RHD calculation is the relatively limited amount of input data required. During the early stages of engine development, when more precise data from measurements are unavailable, this approach makes sense, but can be limited in its scope. Calculations with Elastic Hydro-Dynamic (EHD) models take into account the same reduction of the viscosity. Additionally, the deflections of the solid components reveal critical conditions for the same engine (Fig. 7), which would be hidden using RHD models. Effectively, the results show significant differences: At low engine speed, high combustion pressure produces asperity contact and increases the friction power loss on the rod bearings; main bearings show improvement; At intermediate and high engine speed, the deflection of the crankshaft leads to edge loading and asperity contacts which increase the friction power loss of the main bearings; rod bearings have better hydrodynamic Figure 6: Stribeck Curve. ω: angular speed, η: dynamic oil viscosity, p: specific load on bearing T+S_6_16 17.10.16 17: 01 Seite 46 Aus der Praxis für die Praxis At the same time the operating number ωη/ p decreases and the operating condition of the bearing approaches or enters the mixed lubrication regime, because lower viscosity leads to a smaller oil film thickness (Figure 6). For the bearings, very low viscosity may bring the risk of local loss of oil film and hard contact with the journal - even seizure can occur if the oil and bearing design is not thoroughly optimized. One method for fine tuning is the analysis of the lubrication film with sophisticated software tools. Often RHD (Rigid Hydro-Dynamic) models show the positive impact of lower viscosity oil on friction power loss (Figure 7), providing acceptable minimum oil film thickness for all engine operating conditions. An advantage of the RHD calculation is the relatively limited amount of input data required. During the early stages of engine development, when more precise data from measurements are unavailable, this approach makes sense, but can be limited in its scope. Calculations with Elastic Hydro-Dynamic (EHD) models take into account the same reduction of the viscosity. Additionally, the deflections of the solid components reveal critical conditions for the same engine (Figure 8), which would be hidden using RHD models. Effectively, the results show significant differences: At low engine speed, high combustion pressure produces asperity contact and increases the friction power loss on the rod bearings; main bearings show improvement. At intermediate and high engine speed, the deflection of the crankshaft leads to edge loading and asperity contacts which increase the friction power loss of the main bearings; rod bearings have better hydrodynamic conditions and contribute to friction power loss reduction. At high to very high engine speed, the vertical oval deformation of the rod bearings’ housing bore increases the friction close to the parting line; the main bearings’ power loss suffers from edge loading often driven by flywheel tumbling. In total, the change to lower viscosity oil may produce an over-all improvement of the engine’s friction power loss, but looking at very thin oils, the asperity contacts in the bearings can behave counterproductive. In any case, the wear behaviour connected to the increased surface contact situations has to be observed carefully. Also, the improvement potential depends on the collective of operating conditions. 3.2 Influence of Bearing Geometry From hydrodynamic calculation it is known that the bearing clearance (the lubrication gap) has a significant influence on the friction power loss of a bearing. In the same engine, the maximum clearance due to production tolerances will produce lower frictional losses than the minimum clearance. Looking at the nominal values of the bearing system, downsizing of the engine means reducing both the width Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 47 conditions and contribute to friction power loss reduction; At high to very high engine speed, the vertical oval deformation of the rod bearings’ housing bore increases the friction close to the parting line; the main bearings’ power loss suffers from edge loading often driven by flywheel tumbling. In total, the change to lower viscosity oil may produce an over-all improvement of the engine’s friction power loss, but looking at very thin oils, the asperity contacts in the bearings can behave counterproductive. In any case, the wear behaviour connected to the increased surface contact situations has to be observed carefully. Also, the improvement potential depends on the collective of operating conditions. Fig. 7: Impact of the oil type and engine speed on viscosity and friction power loss of main bearings and rod bearings. EHD sensitivity study, Full Load conditions, 2.0L I4 Gasoline engine 3.2 Influence of Bearing Geometry From hydro-dynamic calculation it is known that the bearing clearance (the lubrication gap) has a significant influence on the friction power loss of a bearing. In the same engine, the maximum clearance due to production tolerances will produce lower frictional losses than the minimum clearance. Looking at the nominal values of the bearing system, downsizing of the engine means reducing both the width as well as the diameters of the journals. The width will influence the size of the bearing area; the diameter affects both area and sliding speed. In a RHD study of hydro-dynamic friction power loss, the operating condition 2500rpm at 50% part load was analyzed (Fig. 8). About a third of the crankshaft bearing friction power loss takes place in the rod bearings. The simultaneous reduction by 10% in journal diameters and effective bearing width reduces the hydrodynamic friction power loss by 26%; a 20% reduction in size reduces the power loss by 47%. By downsizing, beside reducing the size of the functional bearing surfaces, the total size of the associated components such as conrod and crankshaft will also be reduced with positive effect on the mass forces, which are an important portion of the engine operating forces. A reduction of these will also decrease the resulting friction forces. Reducing both journal width and diameter requires intense design work on the complete base engine. In comparison, it seems much easier to reduce only the bearing width; so the crankshaft and the engine block’s housing bores remain unchanged and can be carried over to the new concept. However, the specific bearing load increases, and the minimum oil film thickness is reduced. EHD calculations show that this can produce asperity contacts between the bearings and the journals, increasing boundary friction which may outweigh the benefit from the hydro-dynamic portion of friction power loss (Fig. 9). Finally, the significant reduction of component size may be good for friction optimization, but the stiffness of the system must not be compromised by these design changes; also any increases in the mechanical stresses that arise must not jeopardize the durability of the components. Fig. 8: Downsizing effect of -10 and -20% rel. base size on hydro-dynamic friction power loss. 1,5l inline 4 cylinder gasoline engine at 2500rpm, 50% part load Figure 7: Impact of the oil type and engine speed on viscosity and friction power loss of main bearings and rod bearings. RHD sensitivity study, Full Load conditions, 2.0L I4 Gasoline engine conditions and contribute to friction power loss reduction; At high to very high engine speed, the vertical oval deformation of the rod bearings’ housing bore increases the friction close to the parting line; the main bearings’ power loss suffers from edge loading often driven by flywheel tumbling. In total, the change to lower viscosity oil may produce an over-all improvement of the engine’s friction power loss, but looking at very thin oils, the asperity contacts in the bearings can behave counterproductive. In any case, the wear behaviour connected to the increased surface contact situations has to be observed carefully. Also, the improvement potential depends on the collective of operating conditions. Fig. 6: Impact of the oil type and engine speed on viscosity and friction power loss of main bearings and rod bearings. RHD sensitivity study, Full Load conditions, 2.0L I4 Gasoline engine 3.2 Influence of Bearing Geometry From hydro-dynamic calculation it is known that the bearing clearance (the lubrication gap) has a significant influence on the friction power loss of a bearing. In the same engine, the maximum clearance due to production tolerances will produce lower frictional losses than the minimum clearance. Looking at the nominal values of the bearing system, downsizing of the engine means reducing both the width as well as the diameters of the journals. The width will influence the size of the bearing area; the diameter affects both area and sliding speed. In a RHD study of hydro-dynamic friction power loss, the operating condition 2500rpm at 50% part load was analyzed (Fig. 8). About a third of the crankshaft bearing friction power loss takes place in the rod bearings. The simultaneous reduction by 10% in journal diameters and effective bearing width reduces the hydrodynamic friction power loss by 26%; a 20% reduction in size reduces the power loss by 47%. By downsizing, beside reducing the size of the functional bearing surfaces, the total size of the associated components such as conrod and crankshaft will also be reduced with positive effect on the mass forces, which are an important portion of the engine operating forces. A reduction of these will also decrease the resulting friction forces. Reducing both journal width and diameter requires intense design work on the complete base engine. In comparison, it seems much easier to reduce only the bearing width; so the crankshaft and the engine block’s housing bores remain unchanged and can be carried over to the new concept. However, the specific bearing load increases, and the minimum oil film thickness is reduced. EHD calculations show that this can produce asperity contacts between the bearings and the journals, increasing boundary friction which may outweigh the benefit from the hydro-dynamic portion of friction power loss (Fig. 9). Finally, the significant reduction of component size may be good for friction optimization, but the stiffness of the system must not be compromised by these design changes; also any increases in the mechanical stresses that arise must not jeopardize the durability of the components. Fig. 8: Downsizing effect of -10 and -20% rel. base size on hydro-dynamic friction power loss. 1,5l inline 4 cylinder gasoline engine at 2500rpm, 50% part load Figure 8: Impact of the oil type and engine speed on viscosity and friction power loss of main bearings and rod bearings. EHD sensitivity study, Full Load conditions, 2.0L I4 Gasoline engine T+S_6_16 17.10.16 17: 01 Seite 47 Aus der Praxis für die Praxis as well as the diameters of the journals. The width will influence the size of the bearing area; the diameter affects both area and sliding speed. In a RHD study of hydrodynamic friction power loss, the operating condition 2500 rpm at 50 % part load was analyzed (Figure 9). About a third of the crankshaft bearing friction power loss takes place in the rod bearings. The simultaneous reduction by 10 % in journal diameters and effective bearing width reduces the hydrodynamic friction power loss by 26 %; a 20 % reduction in size reduces the power loss by 47 %. By downsizing, beside reducing the size of the functional bearing surfaces, the total size of the associated components such as conrod and crankshaft will also be reduced with positive effect on the mass forces, which are an important portion of the engine operating forces. A reduction of these will also decrease the resulting friction forces. Reducing both journal width and diameter requires intense design work on the complete base engine. In comparison, it seems much easier to reduce only the bearing width; so the crankshaft and the engine block’s housing bores remain unchanged and can be carried over to the new concept. However, the specific bearing load increases, and the minimum oil film thickness is reduced. EHD calculations show that this can produce asperity contacts between the bearings and the journals, increasing boundary friction which may outweigh the benefit from the hydrodynamic portion of friction power loss (Figure 10). Finally, the significant reduction of component size may be good for friction optimization, but the stiffness of the system must not be compromised by these design changes; also any increases in the mechanical stresses that arise must not jeopardize the durability of the components. 4. Oil Flow 4.1 Bearing Oil Flow The bearings of combustion engines are connected to the pressurized oil circuit [3], which also supplies lubricating oil to systems such as the cylinder head, turbo charger, and the oil jets for piston cooling and small end bushing lubrication. The oil pump is usually driven mechanically by the crankshaft, so the pump rotation speed is proportional to engine speed. In most cases, this means the oil pressure is also approximately proportional to engine speed. As different subsystems within the engine require the delivery of oil at different flow rates, the oil flow needs to be throttled at various positions to maintain the correct oil pressure inside the main gallery. Filtered oil is delivered to the main oil gallery, and the bearings’ share of the total oil flow is distributed to the main bearings. Usually the oil is fed into the upper shell, entering a circumferential groove at the inner diameter which allows the oil to pass into the oilways drilled through the crankshaft, to be available for the lubrication of the conrod bearings. 48 Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 conditions and contribute to friction power loss reduction; At high to very high engine speed, the vertical oval deformation of the rod bearings’ housing bore increases the friction close to the parting line; the main bearings’ power loss suffers from edge loading often driven by flywheel tumbling. In total, the change to lower viscosity oil may produce an over-all improvement of the engine’s friction power loss, but looking at very thin oils, the asperity contacts in the bearings can behave counterproductive. In any case, the wear behaviour connected to the increased surface contact situations has to be observed carefully. Also, the improvement potential depends on the collective of operating conditions. Fig. 6: Impact of the oil type and engine speed on viscosity and friction power loss of main bearings and rod bearings. RHD sensitivity study, Full Load conditions, 2.0L I4 Gasoline engine Fig. 7: Impact of the oil type and engine speed on viscosity and friction power loss of main bearings and rod bearings. EHD sensitivity study, Full Load conditions, 2.0L I4 Gasoline engine 3.2 Influence of Bearing Geometry From hydro-dynamic calculation it is known that the bearing clearance (the lubrication gap) has a significant influence on the friction power loss of a bearing. In the same engine, the maximum clearance due to production tolerances will produce lower frictional losses than the minimum clearance. Looking at the nominal values of the bearing system, downsizing of the engine means reducing both the width as well as the diameters of the journals. The width will influence the size of the bearing area; the diameter affects both area and sliding speed. In a RHD study of hydro-dynamic friction power loss, the operating condition 2500rpm at 50% part load was analyzed (Fig. 8). About a third of the crankshaft bearing friction power loss takes place in the rod bearings. The simultaneous reduction by 10% in journal diameters and effective bearing width reduces the hydrodynamic friction power loss by 26%; a 20% reduction in size reduces the power loss by 47%. By downsizing, beside reducing the size of the functional bearing surfaces, the total size of the associated components such as conrod and crankshaft will also be reduced with positive effect on the mass forces, which are an important portion of the engine operating forces. A reduction of these will also decrease the resulting friction forces. Reducing both journal width and diameter requires intense design work on the complete base engine. In comparison, it seems much easier to reduce only the bearing width; so the crankshaft and the engine block’s housing bores remain unchanged and can be carried over to the new concept. However, the specific bearing load increases, and the minimum oil film thickness is reduced. EHD calculations show that this can produce asperity contacts between the bearings and the journals, increasing boundary friction which may outweigh the benefit from the hydro-dynamic portion of friction power loss (Fig. 9). Finally, the significant reduction of component size may be good for friction optimization, but the stiffness of the system must not be compromised by these design changes; also any increases in the mechanical stresses that arise must not jeopardize the durability of the components. Fig. 8: Downsizing effect of -10 and -20% rel. base size on hydro-dynamic friction power loss. 1,5l inline 4 cylinder gasoline engine at 2500rpm, 50% part load Figure 9: Downsizing effect of -10 and -20 % rel. base size on hydrodynamic friction power loss. 1,5l inline 4 cylinder gasoline engine at 2500 rpm, 50 % part load 4. Oil Flow 4.1 Bearing Oil Flow The bearings of combustion engines are connected to the pressurized oil circuit [3], which also supplies lubricating oil to systems such as the cylinder head, turbo charger, and the oil jets for piston cooling and small end bushing lubrication. The oil pump is usually driven mechanically by the crankshaft, so the pump rotation speed is proportional to engine speed. In most cases, this means the oil pressure is also approximately proportional to engine speed. As different subsystems within the engine require the delivery of oil at different flow rates, the oil flow needs to be throttled at various positions to maintain the correct oil pressure inside the main gallery. Filtered oil is delivered to the main oil gallery, and the bearings’ share of the total oil flow is distributed to the main bearings. Usually the oil is fed into the upper shell, entering a circumferential groove at the inner diameter which allows the oil to pass into the oilways drilled through the crankshaft, to be available for the lubrication of the conrod bearings. In some engines, mostly heavy duty applications, there is an oil supply through the conrod from the big end to the small end bushing; this oil supply is intermittent and reduced to the time during the cycle when overlap occurs between the outlet oil hole in the rod journal and the bearing’s connection to the rod. Optimization of the bearings’ oil flow can be used either to improve the lubrication conditions for other engine sub-systems or to reduce the total oil flow volume and thereby the required oil pump size and power absorption, which accounts for part of the engine friction power loss. Through this effect the bearings also contribute indirectly to engine efficiency. From fundamental rig tests it is known that up to 35% of the oil flow to the main and rod bearings is lost by leakage without major impact on the bearing performance. This gives room for oil flow optimization. The oil flow through a bearing is a system property. It can be described by the Reynolds equation, which consists of three terms for incompressible fluid flow; i.e. Poiseuille’s pressure flow, Couette’s flow due to surface velocity and the squeeze effect: In general, the pressure flow is directly proportional to the pressure difference and the third power of the gap height, and indirectly proportional to the axial width and the viscosity. The gap height depends on the clearance, which is defined as the difference of bearing ID and journal OD in the direction of combustion force. A reasonably small clearance is desirable to promote favorable lubrication behavior but could result in too low an oil flow for adequate cooling at high speed, therefore the use of an eccentric bearing wall shape - perpendicular to the load direction permits a sufficient oil flow to be achieved while maintaining a small clearance. The nominal size of the components influences the flow performance; smaller journal diameters lead to a reduced bearing side face area and thereby to a smaller cross section for the oil flow. For each application, a hydrodynamic calculation must be performed in order to establish the oil flow behavior for the complete engine set of bearings. This requires detailed information such as the system component geometry, temperatures, operating conditions, oil properties and feed pressures. 4.2 Bearing Features to Reduce Oil Flow The geometry of the lubricant gap is influenced by design features on the bearings, which are necessary for production and assembly reasons. These are the parting line ID chamfer, the crush relief and the locating lips in both main and conrod bearings as well as the grooves in mains. The parting face between two half shells is typically machined by broaching. To avoid burrs, which could make contact with the journal and break loose, a small chamfer is needed at the ID side of the parting face. Typical chamfer or edge break size is 0.2 to 0.4mm. The chamfers of the upper and lower shells combine when assembled, producing a triangular section oil channel in the axial direction over the width of the bearings, completely inside the crush relief. The crush relief is an intentional reduction of the bearing wall thickness at the ID just below both the parting lines. It prevents sharp edges and hard contact with the journal in the event of any small radial offset between the two halves of the housing bore, caused by small imperfections in the cap locating process during assembly. The crush relief usually has a depth of 0.01 to 0.03mm. Its circumferential extension measures between 2 and 10mm and depends on the application; whether automotive or heavy duty and using angular split conrod or straight split main bearings. Figure 10: Friction optimization by variation of bearing width (-14 %) needs to consider opposite effects of hydrodynamic and asperity friction power loss T+S_6_16 17.10.16 17: 01 Seite 48 Fig. 9: Friction optimization by variation of bearing width (-14%) needs to consider opposite effects of hydro-dynamic and asperity friction power loss 4. Oil Flow 4.1 Bearing Oil Flow The bearings of combustion engines are connected to the pressurized oil circuit [3], which also supplies lubricating oil to systems such as the cylinder head, turbo charger, and the oil jets for piston cooling and small end bushing lubrication. The oil pump is usually driven mechanically by the crankshaft, so the pump rotation speed is proportional to engine speed. In most cases, this means the oil pressure is also approximately proportional to engine speed. As different subsystems within the engine require the delivery of oil at different flow rates, the oil flow needs to be throttled at various positions to maintain the correct oil pressure inside the main gallery. Filtered oil is delivered to the main oil gallery, and the bearings’ share of the total oil flow is distributed to the main bearings. Usually the oil is fed into the upper shell, entering a circumferential groove at the inner diameter which allows the oil to pass into the oilways drilled through the crankshaft, to be available for the lubrication of the conrod bearings. In some engines, mostly heavy duty applications, there is an oil supply through the conrod from the big end to the small end bushing; this oil supply is intermittent and reduced to the time during the cycle when overlap occurs between the outlet oil hole in the rod journal and the bearing’s connection to the rod. Optimization of the bearings’ oil flow can be used either to improve the lubrication conditions for other engine sub-systems or to reduce the total oil flow volume and thereby the required oil pump size and power absorption, which accounts for part of the engine friction power loss. Through this effect the bearings also contribute indirectly to engine efficiency. From fundamental rig tests it is known that up to 35% of the oil flow to the main and rod bearings is lost by leakage without major impact on the bearing performance. This gives room for oil flow optimization. The oil flow through a bearing is a system property. It can be described by the Reynolds equation, which consists of three terms for incompressible fluid flow; i.e. Poiseuille’s pressure flow, Couette’s flow due to A reasonably small clearance is desirable to promote favorable lubrication behavior but could result in too low an oil flow for adequate cooling at high speed, therefore the use of an eccentric bearing wall shape - perpendicular to the load direction permits a sufficient oil flow to be achieved while maintaining a small clearance. The nominal size of the components influences the flow performance; smaller journal diameters lead to a reduced bearing side face area and thereby to a smaller cross section for the oil flow. For each application, a hydrodynamic calculation must be performed in order to establish the oil flow behavior for the complete engine set of bearings. This requires detailed information such as the system component geometry, temperatures, operating conditions, oil properties and feed pressures. 4.2 Bearing Features to Reduce Oil Flow The geometry of the lubricant gap is influenced by design features on the bearings, which are necessary for production and assembly reasons. These are the parting line ID chamfer, the crush relief and the locating lips in both main and conrod bearings as well as the grooves in mains. The parting face between two half shells is typically machined by broaching. To avoid burrs, which could make contact with the journal and break loose, a small chamfer is needed at the ID side of the parting face. Typical chamfer or edge break size is 0.2 to 0.4mm. The chamfers of the upper and lower shells combine when assembled, producing a triangular section oil channel in the axial direction over the width of the bearings, completely inside the crush relief. The crush relief is an intentional reduction of the bearing wall thickness at the ID just below both the parting lines. It prevents sharp edges and hard contact with the journal in the event of any small radial offset between the two halves of the housing bore, caused by small imperfections in the cap locating process during assembly. The crush relief usually has a depth of 0.01 to 0.03mm. Its circumferential extension measures between 2 and 10mm and depends on the application; whether automotive or heavy duty and using angular split conrod or straight split main bearings. Aus der Praxis für die Praxis In some engines, mostly heavy duty applications, there is an oil supply through the conrod from the big end to the small end bushing; this oil supply is intermittent and reduced to the time during the cycle when overlap occurs between the outlet oil hole in the rod journal and the bearing’s connection to the rod. Optimization of the bearings’ oil flow can be used either to improve the lubrication conditions for other engine subsystems or to reduce the total oil flow volume and thereby the required oil pump size and power absorption, which accounts for part of the engine friction power loss. Through this effect the bearings also contribute indirectly to engine efficiency. From fundamental rig tests it is known that up to 35 % of the oil flow to the main and rod bearings is lost by leakage without major impact on the bearing performance. This gives room for oil flow optimization. The oil flow through a bearing is a system property. It can be described by the Reynolds equation, which consists of three terms for incompressible fluid flow; i. e. Poiseuille’s pressure flow, Couette’s flow due to surface velocity and the squeeze effect: In general, the pressure flow is directly proportional to the pressure difference and the third power of the gap height, and indirectly proportional to the axial width and the viscosity. The gap height depends on the clearance, which is defined as the difference of bearing ID and journal OD in the direction of combustion force. A reasonably small clearance is desirable to promote favorable lubrication behavior but could result in too low an oil flow for adequate cooling at high speed, therefore the use of an eccentric bearing wall shape - perpendicular to the load direction - permits a sufficient oil flow to be achieved while maintaining a small clearance. The nominal size of the components influences the flow performance; smaller journal diameters lead to a reduced bearing side face area and thereby to a smaller cross section for the oil flow. For each application, a hydrodynamic calculation must be performed in order to establish the oil flow behavior for the complete engine set of bearings. This requires detailed information such as the system component geometry, temperatures, operating conditions, oil properties and feed pressures. 4.2 Bearing Features to Reduce Oil Flow The geometry of the lubricant gap is influenced by design features on the bearings, which are necessary for production and assembly reasons. These are the parting line ID chamfer, the crush relief and the locating lips in both main and conrod bearings as well as the grooves in mains. The parting face between two half shells is typically machined by broaching. To avoid burrs, which could make contact with the journal and break loose, a small chamfer is needed at the ID side of the parting face. Typical chamfer or edge break size is 0.2 to 0.4 mm. The chamfers of the upper and lower shells combine when assembled, producing a triangular section oil channel in the axial direction over the width of the bearings, completely inside the crush relief. The crush relief is an intentional reduction of the bearing wall thickness at the ID just below both the parting lines. It prevents sharp edges and hard contact with the journal in the event of any small radial offset between the two halves of the housing bore, caused by small imperfections in the cap locating process during assembly. The crush relief usually has a depth of 0.01 to 0.03 mm. Its circumferential extension measures between 2 and 10 mm and depends on the application; whether automotive or heavy duty and using angular split conrod or straight split main bearings. The locating lips are required to easily and accurately position the bearings axially inside the housing bore, especially if the assembly is done manually. The standard production method for a locating lip is a punching operation from the ID towards the OD, which leaves a depression inside the crush relief area. It is very large compared to the size of the clearance, the crush relief and the ID chamfers. Its depth, width and circular length is in the range of millimetres. To reduce the oil flow, the punched locating lips can be replaced by staked lips (Figure 11). In this concept, the lip is produced by pushing tangentially with a tool on the OD half of the parting surface, which squeezes steel back material and forms a lip at the OD slightly below the par- Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 49 The locating lips are required to easily and accurately position the bearings axially inside the housing bore, especially if the assembly is done manually. The standard production method for a locating lip is a punching operation from the ID towards the OD, which leaves a depression inside the crush relief area. It is very large compared to the size of the clearance, the crush relief and the ID chamfers. Its depth, width and circular length is in the range of millimetres. To reduce the oil flow, the punched locating lips can be replaced by staked lips (Fig. 10). In this concept, the lip is produced by pushing tangentially with a tool on the OD half of the parting surface, which squeezes steel back material and forms a lip at the OD slightly below the parting line. This lip still allows precise assembly but no longer affects the ID hydrodynamic behaviour or oil flow. In main bearings, the grooves and especially their circumferential length determine the size of the pressurized area of the oil feed and hence the oil flow. Most engines use grooves in the upper bearing shells to deliver oil to the crankshaft drillings. Some engines even use partial grooves in the lower main bearings to ensure a continuous delivery to the crankshaft oilways, so the combined groove length in upper and lower shells can clearly exceed 180°. The size of the crush relief and the ID chamfer are even more important for the oil flow in main bearings, if the oil groove is connected to them directly. Shortening of the groove, for example to 150°, can disconnect these features from the pressurized oil supply, clearly reducing the rate of oil drainage from the main bearings (Fig. 11). On the other hand, this reduces the oil supply to the rod bearings somewhat, by shortening the communication period with the crankshaft oil drillings, and leads to intermittent pressure input at the shaft surface. However, the reduced groove lengths are currently state of the art design practice and in widespread mass production. Fig. 11: Oil groove in main bearings with and without connection to the crush relief and the ID parting line chamfer 4.3 Crankshaft Features to Reduce Oil Flow Extending the philosophy of groove length reduction leads to the idea of groove elimination. This makes it necessary to reconsider the crankshaft oil drilling concept (Fig. 12). It leads to a switch from a common design where all upper main bearings have grooves to a diversification with two different upper shell designs with and without oil grooves (Fig. 13). Eliminating the groove means that the main bearings concerned only have oil slots for their own oil supply, connected to the main oil gallery. The other main bearings with oil grooves then each have to pass the oil for two conrod bearings, which is the so called 1MB®2RB oil supply. In comparison to the groove length reduction (>200°®150°, Fig. 11), the elimination of grooves turns out to be even more effective, as calculations for a 6 cylinder inline heavy duty engine show (Fig. 14). Here, the results for oil groove length reduction with a 1MB®1RB feed concept with 7 grooves in the upper main bearings are shown. By changing to a full 1MB®2RB oil feed concept, 4 grooves can be eliminated, which shows the best result for oil flow reduction. Downsizing also shows a significant contribution to the reduction of main bearing oil flow. Beside the size, engine speed also has a significant influence on oil flow. Calculations for an inline 4 cylinder engine show that the savings of oil flow grow with increasing engine speed (Fig. 15). Figure 11: Features on bearings influencing the oil flow: locating lips, crush relief and ID edge break T+S_6_16 17.10.16 17: 01 Seite 49 Aus der Praxis für die Praxis ting line. This lip still allows precise assembly but no longer affects the ID hydrodynamic behaviour or oil flow. In main bearings, the grooves and especially their circumferential length determine the size of the pressurized area of the oil feed and hence the oil flow. Most engines use grooves in the upper bearing shells to deliver oil to the crankshaft drillings. Some engines even use partial grooves in the lower main bearings to ensure a continuous delivery to the crankshaft oilways, so the combined groove length in upper and lower shells can clearly exceed 180°. The size of the crush relief and the ID chamfer are even more important for the oil flow in main bearings, if the oil groove is connected to them directly. Shortening of the groove, for example to 150°, can disconnect these features from the pressurized oil supply, clearly reducing the rate of oil drainage from the main bearings (Figure 12). On the other hand, this reduces the oil supply to the rod bearings somewhat, by shortening the communication period with the crankshaft oil drillings, and leads to intermittent pressure input at the shaft surface. However, the reduced groove lengths are currently state of the art design practice and in widespread mass production. 4.3 Crankshaft Features to Reduce Oil Flow Extending the philosophy of groove length reduction leads to the idea of groove elimination. This makes it necessary to reconsider the crankshaft oil drilling concept (Figure 13). It leads to a switch from a common design where all upper main bearings have grooves to a diversification with two different upper shell designs with and without oil grooves (Figure 14). Eliminating the groove means that the main bearings concerned only have oil slots for their own oil supply, connected to the main oil gallery. The other main bearings with oil grooves then each have to pass the oil for two conrod bearings, which is the so called 1MB →2RB oil supply. In comparison to the groove length reduction (> 200°→150°, Figure 12), the elimination of grooves turns out to be even more effective, as calculations for a 6 cylinder inline heavy duty engine show (Figure 15). Here, the results for oil groove 50 Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 The locating lips are required to easily and accurately position the bearings axially inside the housing bore, especially if the assembly is done manually. The standard production method for a locating lip is a punching operation from the ID towards the OD, which leaves a depression inside the crush relief area. It is very large compared to the size of the clearance, the crush relief and the ID chamfers. Its depth, width and circular length is in the range of millimetres. To reduce the oil flow, the punched locating lips can be replaced by staked lips (Fig. 10). In this concept, the lip is produced by pushing tangentially with a tool on the OD half of the parting surface, which squeezes steel back material and forms a lip at the OD slightly below the parting line. This lip still allows precise assembly but no longer affects the ID hydrodynamic behaviour or oil flow. In main bearings, the grooves and especially their circumferential length determine the size of the pressurized area of the oil feed and hence the oil flow. Most engines use grooves in the upper bearing shells to deliver oil to the crankshaft drillings. Some engines even use partial grooves in the lower main bearings to ensure a continuous delivery to the crankshaft oilways, so the combined groove length in upper and lower shells can clearly exceed 180°. The size of the crush relief and the ID chamfer are even more important for the oil flow in main bearings, if the oil groove is connected to them directly. Shortening of the groove, for example to 150°, can disconnect these features from the pressurized oil supply, clearly reducing the rate of oil drainage from the main bearings (Fig. 11). On the other hand, this reduces the oil supply to the rod bearings somewhat, by shortening the communication period with the crankshaft oil drillings, and leads to intermittent pressure input at the shaft surface. However, the reduced groove lengths are currently state of the art design practice and in widespread mass production. Fig. 10: Features on bearings influencing the oil flow: crush relief and ID edge break 4.3 Crankshaft Features to Reduce Oil Flow Extending the philosophy of groove length reduction leads to the idea of groove elimination. This makes it necessary to reconsider the crankshaft oil drilling concept (Fig. 12). It leads to a switch from a common design where all upper main bearings have grooves to a diversification with two different upper shell designs with and without oil grooves (Fig. 13). Eliminating the groove means that the main bearings concerned only have oil slots for their own oil supply, connected to the main oil gallery. The other main bearings with oil grooves then each have to pass the oil for two conrod bearings, which is the so called 1MB®2RB oil supply. In comparison to the groove length reduction (>200°®150°, Fig. 11), the elimination of grooves turns out to be even more effective, as calculations for a 6 cylinder inline heavy duty engine show (Fig. 14). Here, the results for oil groove length reduction with a 1MB®1RB feed concept with 7 grooves in the upper main bearings are shown. By changing to a full 1MB®2RB oil feed concept, 4 grooves can be eliminated, which shows the best result for oil flow reduction. Downsizing also shows a significant contribution to the reduction of main bearing oil flow. Beside the size, engine speed also has a significant influence on oil flow. Calculations for an inline 4 cylinder engine show that the savings of oil flow grow with increasing engine speed (Fig. 15). Figure 12: Oil groove in main bearings with and without connection to the crush relief and the ID parting line chamfer Fig. 13: Design for main bearings used for a concept of 1 Main Bearing feeding 2 Rod Bearings with diversification of upper main bearings. Fig. 14: Sensitivity study on main bearing oil flow taking into account the component size (bearing ID and width), oil groove length and oil feed strategy from main to rod bearings (1400rpm, 80% of max. peak cylinder pressure, const. power output). Fig. 15: Effect of engine speed and groove elimination from upper main bearings by new crankshaft architecture in an inline 4 cylinder engine. Switch from 1 main feeding 1 rod bearing to 1 main feeding 2 rod bearings Conclusions Higher combustion pressures and additional functionality, such as start stop, in combination with downsizing increase the loads on bushings and bearings. High strength materials based on Aland Cualloys with and without coatings are available to cover the demand for increased durability and wear performance in each different application. Sliding bearings contribute to parasitic losses directly by hydrodynamic friction and indirectly through needing a pressurized oil supply. The direct share can be reduced by a smaller bearing diameter and reduced width. Careful use of lower oil viscosity also reduces their hydrodynamic friction loss. The indirect share can be improved by design features that reduce oil leakage, such as the staked lip, reduced oil groove length in main bearings, and the elimination of main bearing oil grooves. The latter also calls for new oil feed concepts in the crankshaft. Optimization of bearing friction power loss requires careful balance, as actions taken to reduce friction may be counter-productive through increased asperity contacts due to lower oil viscosity or reduced stiffness of the bearing systems. Sophisticated software tools are used to analyze the consequences of design changes for a wide spectrum of operating conditions and to ensure the robustness of the engine bearing system. References [1] Orlowski, K.; Ritterskamp C.; Dohmen J.; Maaßen F.: Advanced Measurement Techniques in the Field of today’s Engine Mechanics 18. Aachen Colloquium Automobile and Engine Technology, 2009 [2] Jückstock, R.: Figure 13: Different oil channel architectures for inline 4 and 6 engines Fig. 12: Different oil channel architectures for inline 4 and 6 engines Fig. 14: Sensitivity study on main bearing oil flow taking into account the component size (bearing ID and width), oil groove length and oil feed strategy from main to rod bearings (1400rpm, 80% of max. peak cylinder pressure, const. power output). Fig. 15: Effect of engine speed and groove elimination from upper main bearings by new crankshaft architecture in an inline 4 cylinder engine. Switch from 1 main feeding 1 rod bearing to 1 main feeding 2 rod bearings Conclusions Higher combustion pressures and additional functionality, such as start stop, in combination with downsizing increase the loads on bushings and bearings. High strength materials based on Aland Cualloys with and without coatings are available to cover the demand for increased durability and wear performance in each different application. Sliding bearings contribute to parasitic losses directly by hydrodynamic friction and indirectly through needing a pressurized oil supply. The direct share can be reduced by a smaller bearing diameter and reduced width. Careful use of lower oil viscosity also reduces their hydrodynamic friction loss. The indirect share can be improved by design features that reduce oil leakage, such as the staked lip, reduced oil groove length in main bearings, and the elimination of main bearing oil grooves. The latter also calls for new oil feed concepts in the crankshaft. Optimization of bearing friction power loss requires careful balance, as actions taken to reduce friction may be counter-productive through increased asperity contacts due to lower oil viscosity or reduced stiffness of the bearing systems. Sophisticated software tools are used to analyze the consequences of design changes for a wide spectrum of operating conditions and to ensure the robustness of the engine bearing system. References [1] Orlowski, K.; Ritterskamp C.; Dohmen J.; Maaßen F.: Advanced Measurement Techniques in the Field of today’s Engine Mechanics 18. Aachen Colloquium Automobile and Engine Technology, 2009 [2] Jückstock, R.: Figure 14: Design for main bearings used for a concept of 1 Main Bearing feeding 2 Rod Bearings with diversification of upper main bearings T+S_6_16 17.10.16 17: 01 Seite 50 Aus der Praxis für die Praxis length reduction with a 1MB→1RB feed concept with 7 grooves in the upper main bearings are shown. By changing to a full 1MB→2RB oil feed concept, 4 grooves can be eliminated, which shows the best result for oil flow reduction. Downsizing also shows a significant contribution to the reduction of main bearing oil flow. Beside the size, engine speed also has a significant influence on oil flow. Calculations for an inline 4 cylinder engine show that the savings of oil flow grow with increasing engine speed (Figure 16). Conclusions Higher combustion pressures and additional functionality, such as start stop, in combination with downsizing increase the loads on bushings and bearings. High strength materials based on Aland Cu-alloys with and without coatings are available to cover the demand for increased durability and wear performance in each different application. Sliding bearings contribute to parasitic losses directly by hydrodynamic friction and indirectly through needing a pressurized oil supply. The direct share can be reduced by a smaller bearing diameter and reduced width. Careful use of lower oil viscosity also reduces their hydrodynamic friction loss. The indirect share can be improved by design features that reduce oil leakage, such as the staked lip, reduced oil groove length in main bearings, and the elimination of main bearing oil grooves. The latter also calls for new oil feed concepts in the crankshaft. Optimization of bearing friction power loss requires careful balance, as actions taken to reduce friction may be counterproductive through increased asperity contacts due to lower oil viscosity or reduced stiffness of the bearing systems. Sophisticated software tools are used to analyze the consequences of design changes for a wide spectrum of operating conditions and to ensure the robustness of the engine bearing system. References [1] Orlowski, K.; Ritterskamp C.; Dohmen J.; Maaßen F.: Advanced Measurement Techniques in the Field of today’s Engine Mechanics, 18. Aachen Colloquium Automobile and Engine Technology, 2009 [2] Jückstock, R.: Perspectives of Friction Reduction on the Base Engine, MTZ, Special Edition: 75Jahre MTZ, 2014 [3] Affenzeller, J., Gläser H.: Lagerung und Schmierung von Verbrennungsmotoren (Die Verbrennungskraftmaschine, N.F., Bd. 8), Wien, New York: Springer, 1996 Tribologie + Schmierungstechnik 63. Jahrgang 6/ 2016 51 Fig. 12: Different oil channel architectures for inline 4 and 6 engines Fig. 15: Effect of engine speed and groove elimination from upper main bearings by new crankshaft architecture in an inline 4 cylinder engine. Switch from 1 main feeding 1 rod bearing to 1 main feeding 2 rod bearings Conclusions Higher combustion pressures and additional functionality, such as start stop, in combination with downsizing increase the loads on bushings and bearings. High strength materials based on Aland Cualloys with and without coatings are available to cover the demand for increased durability and wear performance in each different application. Sliding bearings contribute to parasitic losses directly by hydrodynamic friction and indirectly through needing a pressurized oil supply. The direct share can be reduced by a smaller bearing diameter and reduced width. Careful use of lower oil viscosity also reduces their hydrodynamic friction loss. The indirect share can be improved by design features that reduce oil leakage, such as the staked lip, reduced oil groove length in main bearings, and the elimination of main bearing oil grooves. The latter also calls for new oil feed concepts in the crankshaft. Optimization of bearing friction power loss requires careful balance, as actions taken to reduce friction may be counter-productive through increased asperity contacts due to lower oil viscosity or reduced stiffness of the bearing systems. Sophisticated software tools are used to analyze the consequences of design changes for a wide spectrum of operating conditions and to ensure the robustness of the engine bearing system. References [1] Orlowski, K.; Ritterskamp C.; Dohmen J.; Maaßen F.: Advanced Measurement Techniques in the Field of today’s Engine Mechanics 18. Aachen Colloquium Automobile and Engine Technology, 2009 [2] Jückstock, R.: Figure 15: Sensitivity study on main bearing oil flow taking into account the component size (bearing ID and width), oil groove length and oil feed strategy from main to rod bearings (1400 rpm, 80 % of max. peak cylinder pressure, const. power output) Fig. 12: Different oil channel architectures for inline 4 and 6 engines Fig. 13: Design for main bearings used for a concept of 1 Main Bearing feeding 2 Rod Bearings with diversification of upper main bearings. Fig. 14: Sensitivity study on main bearing oil flow taking into account the component size (bearing ID and width), oil groove length and oil feed strategy from main to rod bearings (1400rpm, 80% of max. peak cylinder pressure, const. power output). Conclusions Higher combustion pressures and additional functionality, such as start stop, in combination with downsizing increase the loads on bushings and bearings. High strength materials based on Aland Cualloys with and without coatings are available to cover the demand for increased durability and wear performance in each different application. Sliding bearings contribute to parasitic losses directly by hydrodynamic friction and indirectly through needing a pressurized oil supply. The direct share can be reduced by a smaller bearing diameter and reduced width. Careful use of lower oil viscosity also reduces their hydrodynamic friction loss. The indirect share can be improved by design features that reduce oil leakage, such as the staked lip, reduced oil groove length in main bearings, and the elimination of main bearing oil grooves. The latter also calls for new oil feed concepts in the crankshaft. Optimization of bearing friction power loss requires careful balance, as actions taken to reduce friction may be counter-productive through increased asperity contacts due to lower oil viscosity or reduced stiffness of the bearing systems. Sophisticated software tools are used to analyze the consequences of design changes for a wide spectrum of operating conditions and to ensure the robustness of the engine bearing system. References [1] Orlowski, K.; Ritterskamp C.; Dohmen J.; Maaßen F.: Advanced Measurement Techniques in the Field of today’s Engine Mechanics 18. Aachen Colloquium Automobile and Engine Technology, 2009 [2] Jückstock, R.: Figure 16: Effect of engine speed and groove elimination from upper main bearings by new crankshaft architecture in an inline 4 cylinder engine. Switch from 1 main feeding 1 rod bearing to 1 main feeding 2 rod bearings T+S_6_16 17.10.16 17: 01 Seite 51