Friction measurements of a piston-ring pack using strain gauges during cold ... geometric dimensions of the piston assembly system are measured using a ...
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Friction measurements of a piston-ring pack using strain gauges during cold start-up of a motorbike engine Anastasios Zavos 1, Pantelis G. Nikolakopoulos 1 1
Machine Design Laboratory, Department of Mechanical Engineering and Aeronautics, University of Patras, Patras, Greece, 26504
Abstract: The ring-pack is the main part of parasitic losses in the engine. In particular, the major contribution of the top compression ring friction is approximately 15-20% of the total friction due to the piston-ring pack. The object of this paper is to measure the piston-ring pack friction during the cold start-up period of a single-cylinder motorbike engine under firing conditions. The basic geometric dimensions of the piston assembly system are measured using a coordinate measuring machine. Strain gauges are used along the cylinder liner and the extracted signals are analyzed using the Labview software. The engine test for friction measurement is taken for a short time of 5 min in order to simulate the cold start-up driving condition. The cylinder temperature is measured using a high-performance thermo camera. A fresh multigrade oil is used during the engine test. The lubricant properties are measured using a capillary tube viscometer from deltalab. In the present study, CFD simulations are carried out in order to predict the friction produced for a new compression ring sliding against a plateau crossed hatched cylinder. The analysis considers a mixed lubrication model combining the Navier-Stokes and Rayleigh-Plesset equations. The contribution of asperity contact in thin lubricant films is obtained. The numerical results are compared with experiments under idle rotational speed. Good agreement is observed between experimental and numerical predictions. Keywords— CFD compression ring, engine cold start-up, friction, strain gauges 1 Introduction Today the automotive industry tries to reduce the CO2 emission levels and improve engines efficiency. Transportation is the major contributory of the global energy consumption. Cars, motor vehicles and trucks are the main sources of CO2 emissions. Low engine speeds during the cold start-up period have shown significant friction losses in city driving. According to the New
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European Driving Cycle, a large portion (60–80%) of the engine emissions occurs in the cold start-up operation. Therefore, many industrial groups have focused to reduce the mechanical losses and investigate ways to improve the performance of tribocontacts at cold start. A large portion 35-45% of the engine friction losses caused from the piston system [1]. The lubricants with the design of the top compression ring have a crucial role on friction, wear and fuel consumption in internal combustion engines [2-4]. Models concluding the friction force and the minimum film thickness in piston rings attracted the attention of many authors due to its importance during the operation of the engines. Yeng [5] proposed an analytical method based on the Reynolds equation solution for a ring-pack system under mixed regime of lubrication. The total friction force and the minimum lubricant film were evaluated, showing the effect of piston ring curvature. Tian [6] presented the dynamic performance of the ring pack tribo-system to determine the wear of ring/liner conjunction and the lubricant film variation. It was found that, both the second ring face worn profile and the piston dynamic motion have a significant role on the top ring tribological behaviour. Dellis [7] also analysed the piston ring–liner pair in partial hydrodynamic regime. The friction force with the use of a single ring test-rig was measured. Two types of piston rings were examined, showing that flatter profile had a lower friction at reversal points. More recently, CFD predictions of the compression ring-cylinder mechanism were presented by Shahmohamadi et al. [8] and Zavos et al. [9]. Flow simulations were performed solving the Navier-Stokes equations. The theory of asperity contact between two rough surfaces by Greenwood and Trip [10] was considered. They found that the numerical results based on the Navier-Stokes approach show more accurate evaluation of viscous friction at low loads and higher piston speeds. The lubrication problem of the piston compression ring was analysed and designed from the tribological point of view, and the results were validated experimentally by many researchers. Furuhama et al. [11], Zavos et al. [12] and Gore et al. [13] presented numerical and experimental results regarding the piston rings-cylinder conjunction. In the study described by Furuhama and Sasaki [11], the total friction was measured using the floating liner method under motored operating conditions. To study the friction performance of the piston-ring pack, the experimental measurements were derived from a small diesel and gasoline engine. Zavos et al. [12] conducted another series of experiments for a four- stroke motor engine using strain gauges under firing conditions. A theoretical and experimental study regarding piston ring-pack tribological
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behaviour was presented. The authors were also examined the influence of the lubricant on the engine performance. They demonstrated that the lubricant with additives can lead to lower friction, as observed from their experimental results. More recently, Gore et al. [13] have shown some potential experimental results for friction on top compression ring under motored conditions. In practical terms, the correlation between experimental and numerical friction to ring–bore was examined and it was obvious that the contact is strictly related with the ring dynamic motion. This paper presents a combined numerical and experimental study for a four stroke motorbike engine. Using strain gauges along the cylinder block, the measured total friction of the ring-pack system was recorded using the program LabVIEW 11. The engine test for friction measurement was taken during the cold start-up period using a fresh multigrade oil. The cylinder temperature was measured using a high performance thermo camera. The experimental results were compared with CFD predictions regarding the top compression ring friction. The basic geometric parameters of the ring-bore system was measured using a coordinate measuring machine. Reasonable agreement was obtained between the numerical predictions and experiment. 2. Overview of the experimental setup 2.1 Test motorbike engine In the current work a single-cylinder four stroke, 107cc, motorbike engine is used due to its small size and air-cooled design. Figure 1 illustrates the basic engine set-up. The piston assembly includes the ring-pack, which consists of two compression rings and a two-land type oil-control rings. The upper compression ring is a barrel faced ring made of chrome plated nitrided steel. The lower compression ring is a taper faced ring with an undercut made of special cast iron. The piston material is a special aluminium-silicon alloy and the cross-hatched honed cylinder bore material is aluminium alloy with high content of silicon. In the benefits of this engine is also included the simple sensors placement with a minimal modification to the external surface of the cylinder block. Additionally, this test engine located on a removable platform and can be
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operated in environmental conditions outside from the laboratory in order to replicate realistic engine conditions.
Fig. 1 Basic set-up of test motorbike engine 2.2 Engine friction and methodology Many studies have been included test rigs or laboratory engines to measure piston-ring pack friction under motorized or fired engine running conditions. The motorised experiments correspond to cold start conditions for low engine speeds according to the New European Emission Driving Cycle [14]. In general, these tests follow some urban driving conditions leading to significant emissions and power losses. In the case of fired conditions, low speed tests with noise have been performed. These performance tests have large number of variables that occur under fired conditions such as engine noise and vibration, thermo-elastic deformations of the cylinder liner and reduced viscosity through temperature deviations. Nevertheless, several investigations showed realistic findings under fired conditions minimizing the other interactions [11-13].
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In order to measure the total friction between the piston assembly and the cylinder liner in mixed/hydrodynamic lubrication regime, a strain gauge was placed on the liner near to the top dead center (TDC), as displayed in Figure 2(a). To prepare the cylinder for directly strains measurement, a section of the cooling fins was removed. To receive more accurate and actual deformations was decided to deepen the area of the strain gauge, so that the thickness of metal between the strain gauge and the piston assembly contact to reduce. The thickness was approximately 4 mm since the deflections was moved closer to the piston assembly friction (shown schematically in Figure 2(b)). This was the only structural modification in order to mount the strain gauge close to piston assembly contact.
Fig. 2 (a) Photo of the strain gauge position, (b) schematic showing engine section including piston assembly contact The sensor system consisted of a low cost, metallic strain gauge of type N11MA512011 with length of 5 mm, nominal resistance of Rg =120 ohm and preinstalled wire. Principally, these gauges have laminated surfaces and integral leads, and were suitable for measurement of both
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static and dynamic strain. It is very important that the strain gauge be properly mounted onto
the test specimen so that the strain is accurately transferred from the test specimen, though the adhesive and strain gauge backing, to the foil itself. Parameter
Value
Unit
Gauge length
5
mm
Gauge factor
2.1
-
Gauge resistance
120
Ω
Sensitivity
0.1%
-
Operating temperature
-30 to 80
ο
C
Table 1. Strain gauge specifications for this investigation
A fundamental parameter of the strain gauge is its sensitivity to strain, expressed quantitatively as the gauge factor (GF). Gauge factor is defined as the ratio of fractional change in electrical resistance to the fractional change in length (strain), GF
R
R.
Therefore, to measure the strain requires accurate measurement of very small changes in resistance. A strain gauge with a gauge factor GF = 2.1 will exhibit a change in electrical resistance of only 0.1%. For a 120 Ω gauge, this is a change of only 0.12 Ω. In Table 1, the main specifications of the foil strain gauge were highlighted. In this study, the amplitude of the reflected deformation signal was governed by the conditions of the piston assembly contact. The export data was performed for a short period of 5 sec. The reflections were captured and analysed with LabVIEW software.
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In order to capture the absolute measures of friction a prior calibration of strain gauge should be made. During the tests, the motorbike engine was reheating for idle engine speed so that local temperature along the sensor position remains constant. Using a high performance thermo-camera along the gauge position and an air-cooling device, the cylinder bore temperature was controlled in a short time of 5 min. After the temperature stabilized, the reflection values were reset due to thermal deformation of the cylinder bore material. Experiments have been shown that the thermal strain of the aluminium cylinder is approximately 23.4 μstrain/oC. Therefore, the setup ensures that reflections on the liner are due to the piston assembly contact.
Fig. 3 Schematic overview of signal processing
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The basic concept of the signals processing is pictured in Figure 3. Regarding the Labview environment, the strain gauge has the following parameters using the first quarter bridge configuration (Type I). This was most commonly used in experimental stress analysis, where ambient temperature is relatively constant. The adjacent resistor in the lower arm was selected to have the same resistance as the strain gauge (R3=Rg). In the present study, solving for the resultant strain of the Quarter-Bridge Type I, it will yield to the following expression:
e 4Vr
1 Rl (1 ) GF (1 2Vr ) Rg
(1)
Error Term
where GF=2.1 is the gauge factor of strain gauge, Rg is the nominal strain gauge resistance, and Rl is the lead-wire resistance. The above strain equation was in error by a factor of the ratio of the lead wire resistance to the nominal gage resistance. This factor was lead in a wire desensitization. Therefore, the error was minimized if the lead wire resistance was small (Rl=0) and/or the nominal gage resistance was large. If ignoring the error term will cause an unacceptable error, then it should be added to the program. To receive the reflection signals, a PC fitted with a data analysis card of the national instruments type NI PCI-6251 with features: 16-Bit, 1 MS / s (Multichannel), 1.25 MS / s (1-Channel) and 16 Analog Inputs was used. A photograph of engine unit and acquisition apparatus is illustrated in Figure 4.
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Fig. 4 Photograph of test engine and computer unit The experimental procedure reported here correspond to the idle cold engine conditions. According to the New European Drive Cycle, these conditions chosen for low speed city driving conditions. It is, therefore, clearly that a strain gauge sensor with controlled temperature conditions along the cylinder bore could provide realistic predictions with good repeatability due to the piston assembly contact. However, one of the most difficult items to realize, is the bonding of the strain gauge to the part whose deformation is desired. Improper bonding can result in an error in measurement due to de-calibration. On the other hand, noise and vibration, temperature and surface topography of piston assembly are key parameters that influence strongly the friction measurements in a fired engine. Near TDC, large combustion pressure through poor lubrication regime was encouraged giving higher noise level on the liner. For this test engine, it was not possible to mount the strain gauge at the top dead centre (TDC) reversal because the position of the cylinder head bolts. So, the gauge was bonded 20 mm from the TDC of piston assembly. This configuration shows slightly reflections of noise in engine cycles with good repeatability during
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engine tests. Furthermore, the influence of temperature is critical for actual measurements. In practical terms, the operation of the strain gauge has a directly relation with temperature during engine tests. To receive real strain gauge reflections, the strain gauge was bonded to the exterior cylinder surface using a high temperature adhesive Vishay BAK200. Simultaneously, the cylinder temperature was controlled to relevant low temperatures due to short period tests. This configuration influence slightly the variation of the lubricant viscosity along the cylinder providing realistic findings. Regarding the surface topography of the piston/cylinder system, new mechanical components (piston, ring-pack and cylinder bore) lubricated with fresh automotive lubricants were used. However, the manufacturing errors of the mechanical parts have a significant role to experimental predictions. 3. Theory 3.1 Compression ring-cylinder conjunction Fully flooded inlet conditions are assumed for a two-phase flow along the ring-bore contact, as well as the load of asperities is considered from Greenwood and Tripp study. In this paper, the generated hydrodynamic pressures into the ring-cylinder gap are obtained using Navier Stokes equations. The momentum equation (1), are coupled with the continuity equation (2) and solved using the finite volume method.
DV V p F Dt
(1)
D V 0 Dt
(2)
where ρ is the lubricant density, V is the fluid velocity vector, p is the static pressure, is the viscous stress tensor and F is the external body forces. The current analysis includes the viscous stress tensor expressed as, T 2 I , where the second term on the right hand
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side is the effect of volume dilation and μ is the dynamic lubricant viscosity. Additionally,
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cavitation originates when in a given temperature, the pressure in the stream of liquid is lower than the pressure of its saturated vapours. The current analysis includes liquid vapour mass transfer, including cavitation, which is solved by the Rayleigh-Plesset volume fraction equation [9]. This simulation includes the in-plane ring load balance in steady state computation. In detail, a sufficient lubricant entrainment into the ring-bore contact is generated as the piston velocity increases. However, at the ends of the piston stroke, thin lubricant films are formed due to high load and slow ring velocity. At each crank angle, the produced load in the ring-cylinder clearance, W = Wh + Wc, is obtained to be equal with the total outward force behind the ring profile, F = Fel + Fbk. Therefore, the in-plane ring motion is summarized as follows: Fel elastic force
Fbk back gas force
Whyd ( ) hydrodynamic generated force
Wcont
(3)
asperities generated force
In the radial ring motion, two outward forces applied between the piston and the back profile of the ring. In practical terms, the total combustion pressure is used to determine the pressure acting behind of the ring, therefore the back gas force is calculated as: Fbk Dcyl Bpbk ( ) . Regarding the ring tension force, the tension force is expressed as, Fel Dcyl Bpel , where the elastic pressure is pel
d gap Er I r D 3 B cyl 2
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for the top ring cross section parameter, I r
BW 3 . Once the 12
hydrodynamic reaction is obtained, the load carrying capacity, Wh, along the ring face is defined as: Whyd Dcyl
B /2
phyd dxdy .
B /2
The lubricant film thickness between the compression ring and the cylinder inner liner is expressed as:
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h( y, t ) hmin t
C B 2
2
y2
(4)
hs y
where hmin(t) is the minimum oil thickness of piston ring-cylinder interface at any time and hs(y) is the artificial parabola shape. The local contact deformation and the thermoelastic contribution are not considered in this analysis.
3.2 Asperity contact model The asperity interactions at the dead centres were taken into account by means of the Greenwood-Tripp model. The asperity height distribution is Gaussian, thus the load Wc carried by the asperities and the real area of contact Ac are expressed by the following equations: Wc
16 2 ' 2 E AF5/2 15
Ac 2 AF2 2
(5) (6)
The statistical function F5/2(λ) is obtained as a relation of the Stribeck oil film parameter (λ = h y, t σ rms
). The probability distribution of asperity heights F5/2(λ), F2(λ) are defined by a fifth-order
polynomial curve as given by ref. [8]. In actual engines, cylinder liners are rough and cross-hatch honed. This topography leads to a non- Gaussian distribution of asperities. Therefore, a more comprehensive wear model that can analyze the wear problem for both Gaussian and nonGaussian surfaces is needed. However, this analysis conforms approximately to a Gaussian distribution of asperities prior to running-in engine condition.
3.3 Temperature and lubricant properties
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The motorbike engine was operated at idle rpm Ω = 1500 rpm. The engine test for friction measurement is taken for a short time of 5 min in order to simulate the cold start-up driving condition. The variation of the cylinder block temperature is presented using data from a highperformance thermo camera type: FLIR SC660 with 640×480 resolution and ±1% accuracy. Figure 6 illustrates the cylinder block temperature distribution under cold start-up conditions.
Fig. 2 Cylinder block temperature distribution under cold start-up engine conditions This motorbike engine is operated with fresh multigrade oil type SAE 10W40. The lubricant properties have been measured using a capillary tube viscometer from deltalab. The basic lubricant specifications are presented in Table 2. To study the lubricating ability of the ring close to the real conditions, the variation of the lubricant viscosity and density is considered in the CFD analysis as given by references [15, 16, 17]. Parameter
Value
Unit
lubricant density, ρο
850.5 @ 40 oC
kg/m3
lubricant viscosity, μο
0.080 @ 40 oC
Pas
specific heat, cp
1985
j /kgK
thermal conductivity, k
0.143
W/mk
ao
1x10-8
m2/N
βο
4x10-2
─
Table 2. Lubricant SAE 10W40 specifications at atmospheric pressure and ambient temperature 40 oC
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4. CFD Solution This simulation model considers the effect of the in-plane compression ring motion in quasistatic equilibrium. The algorithm is updated at any crankshaft position by considering the piston speed, the minimum lubricant film, the combustion pressure and the asperity load. External and internal forces acting in the radial direction, influence the tribological behaviour of the compression ring. Figure 3 shows the input boundary conditions on the top compression ring.
Fig. 3 CFD boundary conditions on the compession ring
In Eq. 7 the variation of the boundary input pressures is represented as follows:
B pin pc Upstroke motion p x, 2 pout pc Downstroke motion B pout pa Upstroke motion p x, 2 pin pa Downstroke motion p p c bk
(7)
The convergence criterion of the load balance was used according to the expression (8). If this criterion is not fulfilled, then the minimum film thickness hmin is changed and the interaction loop of lubricant thickness (9) is repeated. The numerical parameter e is considered e = 0.05 for this analysis.
X
W ( )-F ( ) max{F ,W }
0.1%
n 1 n hmin (1 e)hmin ,n 1
(8)
(9)
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Generally, computational fluid dynamic (CFD) analysis offer flexibility to modify design parameters and to associate complex conditions to flow and structural studies. According to the relevant published works [8, 9], CFD predictions have shown a better agreement with the experimental results in relation with those from the Reynolds equation solution under fullhydrodynamic lubrication regime. The basic analysis is almost the same as the work of Zavos and Nikolakopoulos [9] except the use of surface texturing on the barrel ring surface. The convergence in the final solution for the mixed hydrodynamic lubrication problem is obtained when both pressure and load balance criteria are attained simultaneously, after 5000 iterations, and then the simulation moves to the new crank angle. For greater accuracy, the convergence tolerance is assumed to be equal 10-6 for all residual terms. 4.1 Numerical ring friction The total friction force along to the top compression ring contributes significantly to the frictional losses of the engine. The main contribution of the top compression ring friction is approximately 15-20% of the total friction which are produced by the piston assembly system. The total engine friction was obtained by comprising two sub-models (boundary/mixed and hydrodynamic regime). The hydrodynamic friction due to viscous shear stress of the lubricant film and the boundary/mixed friction due to asperity contact load calculated by Eq. (10) for this analysis.
h ftotal p V ( A Ac ) o Ac aspWc 2 h
(10)
boundary / mixed friction viscous friction
It should be noted that the shear stress through lubricant film at hydrodynamic conditions is
h defined as p V . Regarding the boundary lubrication regime, the non-Newtonian 2 h Eyring shear stress and the boundary shear strength of the sliding surfaces are considered
o 2 106 Pa and μasp = 0.2. In addition, it should be referred that the effects of coatings, such as the surface topography including the material mechanical properties, could be predicted for a boundary/mixed friction model. If more factors need to be studied, the corresponding surface topographical parameters should be adopted using an atomic force microscope (AFM) in contact.
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5. Results and Discussion In this work, CFD predictions are carried out with a barrel faced compression ring and a cylinder surface model of a four stroke motorbike engine. The first compression ring is a chromium plated steel ring characterized by radial width W = 3.35 mm and a thickness of B = 0.5 mm. The top ring surface roughness is σr = 0.3 μm and the bore surface roughness is σcyl = 0.15 μm as obtained from surface topography measurements. The cylinder bore is assumed to be a complete circular cylinder with nominal diameter Dcyl = 52 mm and the piston-ring end gap is taken as dgap = 0.25 mm. The chosen lubricant is a fresh multigrade SAE 10W40 and the average temperature of the lubricant is assumed to be equal with cylinder block temperature. Due to the small stroke distance, the average temperature is assumed to be Tav=34 oC. For this motorbike engine investigation, Top Dead Centre (TDC) is defined at the crank angle 360o and the maximum combustion pressure value considered at the crank angle φ = 380ο, which is after TDC position. Figure 4 (a-b) shows the variation of the combustion pressure and the piston linear velocity at idle rotational speed 1500 rpm.
Fig. 4 (a) Combustion pressure versus crank angle (b) Piston linear velocity versus crank angle Figure 5 (a) shows the variation to the measured friction force versus the crank angle using strain gauges. In detail, these experimental results performed for idle rotational speed 1500 rpm and cold start-up conditions. Piston assembly group is the main source of friction in an engine (up to 40%) with losses due to ring-pack being the greater [1]. Generally, the major contribution of the compression ring friction is approximately 15-20% of the total piston assembly friction
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[11,13]. Figure 5(b) presents the numerical ring friction for a barrel-faced ring with curvature height 3 μm to an oval bore conjunction during cold start-up under the same conditions as those in Figure 5(a). Good agreement can be seen between the results. It can be observed that the boundary friction promotes the total friction at the vicinity of TDC during the power stroke. Note that the corresponding lubricant film is quite thin. Therefore, the prevalent regime of lubrication is mixed, with thin films encountered asperity interactions. As the piston velocity becomes larger, the Couette flow pronounces the ring-bore gap, forming thicker films and lower friction. However, there are some differences between the results. In fact, the contribution of asperities between the ring and cylinder bore is essential in the experiments predictions. At the same time, the non-ideal ring and cylinder profiles plays a vital role to the prediction of the total friction due to reduced clearance. These data are not taken into account in the present analysis.
Fig. 5 Comparison of measured and numerical top ring friction
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3. Conclusion In the current paper, a combined experimental and numerical study of the piston assembly system has been presented. The basic geometry of compression ring- cylinder tribosystem was obtained from a real four stoke motorbike engine. The ring-pack friction was measured at idle rotational speed during cold start-up using strain gauges. The CFD analysis simulates in plane ring boundary/mixed isothermal lubrication regime. The interactions and the rheological parameters variation along the ring width were taken into account. The numerical results were verified against experimental results, showing a sensible agreement. The present simulation model shows that more work was needed to predict the tribological characteristics of the top compression ring. Gas blow-by, bore out-of-roundness and ring face profile are very important parameters on the ring-liner and ring-groove cooperation. Additionally, the ring motion with elastic deformations into the piston groove has a dominant influence in the engine operation. Thus, an elasto-dynamics model would provide a more complete analysis for the operating conditions of interest. Acknowledgement This work was supported by Grant Ε.039 from the Research Committee of the University of Patras (Programme K. Karatheodori). References: 1. Holmberg, K. Andersson, P. Erdemir, A. Global energy consumption due to friction in passenger cars, Tribol. Int., 2012, 47, 221–34
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2. Senatore, A. and Aleksendric, D. Advances in piston rings modelling and design, Recent Patents on Engineering, 2013, 7(1), 51-67
3. Zavos, A.B. and Nikolakopoulos, P.G. Simulation of piston ring tribology with surface texturing for internal combustion engines, Lubrication Science, 2015, 27, 151–176
4. Zavos, A. and Nikolakopoulos, P.G. A tribological study of piston rings lubricated by power law oils, Proc. Inst. Mech. Eng. J: J. Eng. Tribol., May 2016, 230(5), 506-524
5. Jeng, Y. Theoretical Analysis of Piston-Ring Lubrication Part 1-Fully Flooded Lubrication, STLE Tribol. Trans., 1992, 35, 696–706
6. Tian, T. Dynamic behaviors of piston rings and their practical impact. Part 2: oil transport, friction and wear of ring/liner interface and the effects of piston and ring dynamics, Proc. Inst. Mech. Eng. J: J. Eng. Tribol., 2002, 216, 229-248
7. Dellis, P.S. Effect of friction force between piston rings and liner: a parametric study of speed, load, temperature, piston-ring curvature, and high-temperature, high-shear viscosity, Proc. Inst. Mech. Eng. J: J. Eng. Tribol., 2010, 224, 411-426
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9. Zavos, A. and Nikolakopoulos, P.G. Cavitation effects on textured compression rings in mixed lubrication, Lubrication Science, 2016, doi:10.1002/ls.1341
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