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Experimental and simulative research advances in the piston assembly of an internal combustion engine
Author’s Accepted Manuscript Experimental and simulative research advances in the piston assembly of an internal combustion engine C. Kirner, J. Halbhuber, B. Uhlig, A. Oliva, S. Graf, G. Wachtmeister www.elsevier.com/locate/jtri
PII: DOI: Reference:
S0301-679X(16)00139-0 http://dx.doi.org/10.1016/j.triboint.2016.03.005 JTRI4111
To appear in: Tribiology International Received date: 22 December 2015 Revised date: 27 February 2016 Accepted date: 5 March 2016 Cite this article as: C. Kirner, J. Halbhuber, B. Uhlig, A. Oliva, S. Graf and G. Wachtmeister, Experimental and simulative research advances in the piston assembly of an internal combustion engine, Tribiology International, http://dx.doi.org/10.1016/j.triboint.2016.03.005 This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting galley proof before it is published in its final citable form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental and simulative research advances in the piston assembly of an internal combustion engine C. Kirnera , J. Halbhubera , B. Uhliga , A. Olivaa , S. Grafa , G. Wachtmeistera a Institute
of Internal Combustion Engines, Technical University of Munich
Abstract The piston assembly is the most complex tribological system within the internal combustion engine. In order to fully exploit its optimization potential a high level of system understanding is required. Therefore, in this paper experimental studies of two single cylinder engines are combinded with CFD-simulation of the piston ring pack. The introduced measurement and simulation techniques enable a holistic approach to the investigation of the tribological conditions of the piston assembly. The results – like the comparison of crank-angle resolved oil film thickness to the piston assembly friction measurement – illustrate the implications of design parameters of the piston assembly on the functional parameters friction, oil consumption, blow-by, wear and acoustics and thereby permit purposive system optimization. Keywords: Piston assembly, Lubricating oil film thickness measurement, Floating liner engine, CFD piston rings 1. Introduction In 2009 the EU member states have committed themselves to reduce the emission of polluting greenhouse gases, especially CO2 , by at least 20 % by 2020 compared to 1990. Therefore emissions of 95 g CO2 /km are the limit for new registrations of motorcars as of 2020 [1, 2]. Compared to the determined average value of 126.7 g CO2 /km by the EEA (European Environment Agency) [3] in 2013, the automotive industry is faced with tremendous economic and technical challenges, as a study of the IKA (Institute for Automotive Engineering, RWTH Aachen) shows [4]. Hence the reduction of CO2 emissions and the related fuel consumption is the key task in the development of modern combustion engines. A technical method to attain these target values, is to improve the mechanical efficiency. According to Holmberg [5] between 4.3 and 7.8 % of the fuel energy is dissipated by the friction between the piston assembly and the cylinder liner. Accordingly, the optimization of the piston assembly provides a good opportunity for the development of improved future engines. To exploit the potential for improvement, a detailed knowledge of the tribological processes is essential. Due to the constantly changing conditions like viscosity, velocity and pressure, the piston assembly is one of the most complex tribological systems inside the internal combustion engine [6]. In addition to the friction, the design of the piston assembly significantly influences the oil consumption and blow-by losses. Furthermore, wear and acoustic effects have to be taken into account, especially with the contact alteration of the piston. The major challenge in optimizing the piston assembly is Preprint submitted to Tribology International
thus to find measures that optimize the friction without influencing other tribological phenomena unfavourably. An evaluation of new design features with regard to effects on the friction-force is not sufficient for an effective improvement of the system piston assembly. In fact, the development of innovative components requires the holistic examination of all influences on this tribological system.
2. Theoretical basics In this paper, effects of a component variation – preload of the oil scraper ring – on the piston assembly friction, blow-by gas flow and the lubricating oil film thickness are investigated. The latter is crucial for friction, oil consumption and wear. So far little research has been done on dependencies between all these phenomena on the piston assembly. In many publications only individual effects, such as the study of friction and wear behavior at a tribology test rig [7] have been investigated. Interactions between the tribological behavior and other effects, such as changes in blow-by that occur during the combustion cycle, have not been analyzed yet. During the strokes the boundary conditions like viscosity, velocity and pressure change and the piston assembly passes states from boundary friction to hydrodynamic friction, as shown in Figure 1. The hydrodynamic friction represents a significant contribution to the friction losses, which is defined by the product of actual friction force and relative speed. The piston stops for a very short period of March 17, 2016
Nomenclature d h n p R t v
distance lubricating oil film thickness engine speed pressure surface roughness time velocity
α η κ µ
volume fraction viscosity piston tilt angle friction coefficient
Film thickness h Friction coefficient µ
time at the top and bottom dead center where the relative speed is zero. In these areas of boundary friction the oil film thickness is almost zero and wear occurs. The range of high relative speed in between the dead centers is of particular interest since the significant losses arise here [8].
A B
C
anti-thrust side bottom dead center crank angle gas exchange top dead center tangential force ignition top dead center indicated mean effective pressure top dead center thrust side
I II III IV
intake compression power exhaust
temperature, viscosity and pressure conditions at the piston rings are different. To understand the different friction states, the knowledge of lubrication oil film thickness is elementary, which – according to Figure 1 – increases with higher speed while viscosity and pressure are maintained. As a consequence, it is sensible to measure and compare friction-force and lubricating oil film thickness between piston and cylinder liner. In [12] the lubricating oil film thickness between the piston ring and the cylinder liner as well as the twist of the piston ring were considered in a modern 1.6 liter gazoline engine. By quantifying the difference in voltage drop between rings, oil film, and liner the research paper [13] presents an oil film thickness measurement method. Paper [14] provides measurement of oil film thickness by using noninvasive ultrasonic sensors on a motocross engine. In [15] crank angle resolved friction-forces were measured on an utility vehicle piston assembly. Simultaneously the lubricating film thickness between the cylinder liner and a piston ring was determined by means of inductive sensors. In [16] the floating-liner principle was used to analyse the running in effect of the cylinder liner. In [17] a piston assembly friction measurement was carried out on a singlecylinder gasoline engine using the IMEP method. The calculation of the friction-forces – as described in [18, 19] – is possible with the knowledge of the crank angle resolved lubricating film thickness between the friction partners. The oil consumption plays an important role in the design of modern internal combustion engines, and it is a subject of current research. In [20] the oil consumption was detected by means of a sulphur tracer and weighing. Moreover the influence of different laser-textures on the cylinder liner was measured. In particular, the reverse blow-by contributes to the oil consumption, which in turn is influenced by ring land pressures and piston ring movement. So in addition to blow-by, piston ring movement and ring land pressure were measured to understand their
A: Boundary friction (h → 0) B: Mixed friction (h ≈ R) C: Hydrodynamic lubrication (h R)
0
ATS BDC CA GTDC FT ITDC IMEP TDC TS
Boundary friction Hydrodynamic friction 0 viscosity η · velocity v pressure p
Fig. 1: Film thickness and Stribeck curve acc. to [8, 9, 10]
The exemplary friction-force plot in Figure 2 illustrates that a separate consideration of the power cycle or other parts of the stroke is not effective in terms of an improvement concerning the tribological conditions. Thus, in order to understand the occurring phenomena on the piston assembly sufficiently, it is important that the complex effects are considered for the entire working cycle. According to [11] the dead centers indicate mixed friction conditions having different characteristics in TDC and BTDC, as liner 2
small influence of the relative ring end gap positions of the three piston rings to each other on the blow-by is detected. However, the position of the ring end gaps to each other has a greater influence on the oil transport. Other investigations are limited to the CFD-simulation of very local phenomena in the ring pack. The relative velocities between rings and cylinder liner during a working cycle and the low film height can cause cavitation effects on the ring running surface. These effects modify the tribological conditions between ring and cylinder liner and can have an impact on all development objectives (blowby, friction, oil transport and wear) in the piston ring pack [24, 25]. New studies focus on the calculation of the flow in the ring pack by means of a 2D-CFD approach where the ring movement is considered. This method has been successfully applied to a diesel and a gasoline engine [26, 27]. The calculation method, the achieved results and also the possibilities of validation of the simulation will be described in the following.
Friction force [N]
Boundary friction Hydrodynamic friction Overall friction
0
90
180
270 360 450 540 Crankangle [◦ CA]
630
720
Fig. 2: Schematic friction force acc. to [19, 11]
3. Method effect on oil consumption. The reduction of friction and blow-by amount is often a conflict of objectives which is the reason for measuring these parameters. In general, decreasing piston clearance results in better acoustics, but higher friction losses. Studies of correlations between friction- and acoustics-behavior were carried out in [21]. However, changes in lubricating oil film thickness and blow-by gas flow were not examined here. Because of its complex modelling, the field of CFD of the ring pack is still relatively young. The 0D- and 1D- simulations that are based on the chamber orifice principle have been established as state of the technology for a long time [22]. The advantages of this method are the simple modeling and short calculation times. However there are also a number of disadvantages. In order to achieve reliable results, the flow coefficients for the small gaps at the piston rings (cylinder liner, ring flanks and ring end gaps) must be specified or rather have to agree with the measured data for the blow-by. Also simplifications for oil transport in the ring pack are made. The mainly used Reynolds equations for thin films are not a valid assumption for calculating the oil transport in the ring pack in the areas behind or between the rings. The CFD simulation provides new knowledge to improve the simplified models in the 0D-/1D-simulation. A 3D CFD-simulation of the gas and oil flow in the ring pack has already been carried out in 2011, see [23]. In this study, the piston rings are considered rigid at certain positions of the ring groove. Depending on the position of the rings, the results of the blow-by and oil transport were significantly different. The ring motion could be identified as an important influencing factor for the transport phenomena in the ring pack, hence the piston ring motion should not be neglected. As a further result, only a
Measurements are carried out using two single-cylinder gasoline engines. Both engines have the same crank drive and cylinder capacity, as the engine data show in Table 1. Piston, piston rings and conrod are series production parts of a gasoline engine with a cylinder capacity of 2.0 liter. The cylinder liner design is close to the production engine. This design ensures that all results are compatible between the research engines and transferable to production engines.
Engine parameter
Value
Cylinder capacity
0.5 L
Bore
82.5 mm
Stroke
92.8 mm
Compression ratio
9.6:1
Operating points
Value
Engine speed
1500 rpm
IMEP
2000 rpm
5 bar
6 bar ◦
Oil temperature
90 C
Coolant temperature
90 ◦C
Table 1: Engine data
To prove the relation between piston assembly friction and oil film thickness, the tangential force of oil control ring is reduced (sec 4.1). This change of design makes it possible to consider the trade-off between friction reduction and oil consumption. The friction measurement is realized by a 3
floating liner engine (built up in 2010), and the measurement of the oil film thickness, ring land pressures, piston secondary, and piston ring movement is applied in a second engine (built up in 2014) at an engine speed of 2000 rpm and an IMEP of 6 bar, respectively (sec 3.1–3.3 and 4.1). The results of the CFD simulation (sec 3.4 and 4.2) are compared to the measurements at an engine speed of 1500 rpm and an IMEP of 5 bar, Table 1.
Probe Cylinder housing
O-ring
Liner Ferrule
Nut 1
3.1. LIF Measurement of lubricating oil film thickness Oil film thicknesses are measured using laser-induced fluorescence (LIF), similar to [28, 29, 30, 31, 32]. A monochromatic laser-light excites fluorescence in the lubricant and the added dye. Optical fibers are used for the transport of the laser- and fluorescent light. 2 a
b Snap ring
d
%
0 450
A prior calibration that allows to put oil film thicknesses in relation to the fluorescent intensity enables absolute measures of lubricant film thicknesses with the help of LIF. Calibration techniques that allow adjustment of lubricant film thicknesses and measurement of fluorescent intensity are shown in [30, 32, 34]. To take the adjustment of the optical setup and the quality of the measurement point on the liner (regarding polishing and assembly-offset) into account, a reference experiment was conducted directly on the experimental engine. Piston rings and precision foils with 12 µm thickness were used. An increasing number of foils (up to 4 with a common thickness of 48 µm) were clamped between the piston ring and the liner, and under the addition of fluorescent oil, a linear relation between the number of sheets and fluorescent intensity could be measured. The setup ensures that reflections on the piston rings are considered within the calibration procedure and the measurement. Furthermore, the influence of oil temperature, like in [30, 31] was investigated. By increasing the oil temperature from 20 ◦ C to 70 ◦ C, the fluorescence intensity decreases to 85 % of the value at 20 ◦ C. With a further raise to 90 ◦ C oil temperature, the fluorescence intensity was reduced to 75 % of the initial value at 20 ◦ C. The oil- and coolant temperature were set to 90 ◦ C in the testing scenario, for this reason the calibration figures for this temperature were applied in the signal processing. In order to measure the lubricating oil film thickness between the piston assembly and the cylinder liner in hydrodynamic lubrication state, two measuring points were placed on the liner, as displayed in Figure 4. Measuring
Exciting Laser Dye Absorbtion Dye Emission Dichroic Filter Emission Filter
50
500 Wavelength [nm]
Piston assembly
e
g f 100
Coolant
Fig. 4: LIF measuring points 1 and 2, piston assembly in vmax acc. to [33]
c c
Optical fiber
550
Fig. 3: Optical setup acc. to [33]
A continuous-wave laser with a wavelength of 473 nm leads to a sufficient excitation, and the emission spectrum is located at higher wavelengths due to the stokes-shift, with a maximum around 500 nm. Figure 3 shows one beam path of the optical setup for sixteen synchronous points of measurement. Laser light from a single laser (a) is divided by a cascade of beam-splitters into sixteen optical paths (b), each of which is coupled by two mirrors (c), a dichroic filter (d), and a fiber coupling (e) into a single fiber leading to the measurement point in the engine. There, laser light exits the fiber directly into the oil film and induces fluorescence that returns through the same fiber. Dichroic filters (d) and additional emission filters (f) separate fluorescent wavelengths from reflections. The fluorescent light is focused on a photodiode (g) and its intensity is measured. 4
point 1 is used for the examination of the lubricating oil film thickness between the piston rings and the cylinder liner at the hydrodynamic lubrication state. The piston rings have a speed of 96–99 % of top speed. Measuring point 2 is positioned in such a way that the illustrated position on the piston skirt is measured at the top speed of the piston. Probes are incorporated inside the cylinder housing and they are fixed with nuts to the cylinder liner. The waterproofing is realized by two radial caulking o-rings for one measuring point; one of them is for sealing between cylinder housing and probe and the other one for in between cylinder liner and probe. Advantages of the radial sealing between the cylinder liner and the probe include close-toproduction liner shapes and surfaces, as well as greatly reduced influence on the liner distortion with proper cooling. The optical fibers are glued in stainless steel ferrules and polished for both a high emission of the laser light and the coupling of the fluorescence light back to the fiber. To fit the ferrules flush with the liner surface, an aluminium cylinder is positioned inside the liner with a clearance fit and heated such that the clearance fit changes into a interference fit. This ensures that the ferrules are glued flush with the liner. Since the ferrules involving the fibers are only glued inside the probes, the dismounting and service of components in this assembly is possible without destroying the optical fibers.
6 FM
3
5
7 2 FF
8
FS 4 1: 2: 3: 4: 5: 6: 7: 8:
Cylinder liner Cooling jacket Load cell Radial bearing Basic engine Cylinder head Radial sealing Piston assembly
1
Fig. 5: Schematic floating liner diagram
is constant, operation point independent and compensated within the measurement data processing. Thereby the friction of the piston assembly can be determined isolatedly. The central design element of a floating liner engine is the sealing between the floating cylinder liner and the cylinder head. The combustion chamber has to be sealed without blocking the axial degree of freedom of the cylinder liner. In this case the cylinder head is extended into the cooling jacket, which has the same bore diameter as the cylinder liner. The radial sealant is realized by a high temperature gasket. With this concept, a gas force compensation, that is required for the mentioned Musashi and PIFFO concept is dispensable [22, 36, 37, 38]. For a precise measuring of the piston assembly friction force it is necessary to control disturbance variables like cylinder pressure and sealing, engine vibrations, crankcase pressure, and temperature deviations. With this test bench, the external conditioning and the employed measuring procedure, it is possible to handle these disturbance variables, and a high level of reproducibility is achieved. The reproducibility is described by the deviation of two measurements between which a simulated change of components is performed. Therefore the engine is disassembled after the first measuring cycle, as it is carried out for a change of components, but assembled with exactly the same components. After conditioning the engine the identical measurement cycle is operated, and the results are compared. To gain results of high reproducibility, a stationary operating point and equal center of combus-
3.2. Crank angle resolved measurement of piston assembly friction To measure the piston assembly friction in an internal combustion engine, different methods like strip method, indication method, IMEP method [35] and the floatingliner-method are available [36, 22]. Solely the floating liner method enables direct crank angle resolved measurements of the piston assembly friction under actual internal combustion engine conditions. The IMEP method, carried out in [17], allows the indirect crank angle resolved measurement of the piston assembly friction. The single cylinder test station is based on the same series engine as the LIF equipped engine whose basic data is summarized in Table 1. There are several design concepts for floating liner systems like the Musashi engine or the PIFFO engine, which are explained in detail at [16, 22, 36, 37, 38]. The basic concept of the second generation floating liner engine is pictured in Figure 5. The assembly of cylinder liner and cooling jacket is mounted on the basic engine by two load cells, which are positioned in the crank shaft axis. This arrangement of the load cells reduces the influence of lateral piston forces on the measurement results. The load cells are prestressed to allow the measurement of compressive and tractive forces. The lateral piston force is retained by a radial bearing. Hence the measuring device is mounted in shunt mode and a calibration is necessary to define the ratio between the applied friction force and the force measured by the load cells. The shunt force ratio 5
tion mass are required. The combustion cycle averaged standard deviation of the measured friction force curves is less than 2.5 N for every operating point of the performance map [11, 39]. This measurement device allows to examine variations of piston assembly components as well as changes of engine boundary conditions and to evaluate the results. 3.3. Setup for axial piston ring movement, piston secondary movement and ring land pressure measurement The measurement of the axial piston ring movement is conducted with the help of eddy current sensors. The position of the sensors is illustrated on the left-hand side of the Figure 6. For the integration of the sensors in the top land two through-holes, through which the sensors can be assembled, are manufactured. Both of them are sited in the ring land between the first and the second compression ring that leads to lower mechanical and thermal stresses for the sensor measuring the motion of the first piston ring. The sensor cables are drawn through radial bores into the piston. Both this radial bores and the through-holes in the top land are pasted up after the mounting. The sensor cables are routed via a mechanical linkage to the measurement data acquisition. On the right-hand side of Figure 6, the position of the resistive pressure sensors for the ring land pressure measurement and two eddy current sensors for recording the piston secondary movement are marked out. The pressure sensors have a diameter of 1.9 mm, and therefore a flush mounting to the outer surface is only possible on the first ring land. In the second ring land, however, no flush mounting is possible, which is why the installed sensors are recessed. One additional eddy current sensor for the detection of the piston secondary movement is installed on the anti-thrust side of the piston, see Figure 11.
Fig. 7: Simulation mesh and computational domain [27]
3.4. CFD-computation of the flow in the piston ring pack Basically, many preliminary considerations are necessary for the numerical calculation of the flow in the ring pack. The computational domain can be defined in different ways depending on the modeling depth. The known options are the 3D full model, a 3D segment and a 2D cut through the ring pack, [27]. The advantages of the 2D cut are the lower number of cells that are necessary for the description of the flow region and the simpler implementation of the ring motion (axial, radial, twist) that is important for the flow. For 3D simulations the complexity of the ring motion increases with a growing number of degrees of freedom. The disadvantage of the 2D simulation is the neglection of the ring end gaps. Despite this simplification, the flow phenomena can be represented well, [26, 27]. Before a computational mesh can be built, the relevant area for the flow has to be derived from CAD data. In addition to the deformation of the cylinder liner resulting from the assembly process, the thermal deformation of the components need to be considered and averaged for the 2D cut. A differentiation between the thrust and the anti-thrust side has not been made. The derived geometry and a computational mesh for the CFD-simulation can be seen in Figure 7. Through the narrow gaps on the ring flanks or rather between the rings and the cylinder liner, locally a fine res-
Eddy current sensors Through-hole
Sensor cables
Resistive pressure sensors
Eddy current sensors
Fig. 6: Axial ring movement (left), piston secondary movement and ring land pressure measurement setup (right) acc. to [33]
6
olution for the mesh has to be chosen. It was ensured that the computational mesh is widely orthogonal structured, which has numerical advantages as well as advantages for the mesh movement. A detailed description of the mesh movement can be found in [26]. The consideration of the ring motion in the CFD simulation guarantees a realistic depiction of the occurring effects. Pressure and temperature of the combustion camber on the one side and pressure in the crankcase on the other side are defined as boundary conditions of the simulation. Moreover, the temperatures at the cylinder liner, the piston and the rings (averaged over one working cycle) are specified. The temperature data may be derived from FEM-simulations or, as in this case, from measurements on the research engine. The motion of the rings is externally calculated in a 1D simulation and then set in the CFD simulation. The movement of the piston is taken into account by a transformation of the coordinate system. The cylinder liner moves from the perspective of the piston-fixed coordinate system with negative piston velocity. The CFD simulation also includes the oil in the piston ring pack. For this exchange models have to be defined between the phases air and oil. The distribution of the oil in the ring pack can be calculated by the initialization of oil in the reservoirs behind the rings. In order to avoid the ring pack running dry the simulation assumes that in the piston downward movement the area below the oil ring is fully filled with oil. The aim of this multi-phase simulation is to calculate the oil distribution in the ring pack or rather the calculation of oil transport phenomena, which has been successfully demonstrated in [27]. The validation of the flow results can be carried out in various ways by the available measurement results from the research engine. The measured ring motion at the engine can be used to validate the ring motion results of the 1D simulation. The ring land pressures represent another important validation opportunity to calculate the correct pressure ratios over the rings. Finally, the integral result of the blow-by from the CFD-simulation can also be validated with the measured blow-by volume flow rate.
κ +
+
-
dT S
piston ring land
dT S
piston skirt
dAT S
coordinate system
piston skirt
κ
ATS
TS
0,3
dT S
piston ring land
d [mm], κ [°]
0,2 0,1
dT S
piston skirt
0 -0,1 -0,2
dAT S
I
κ piston skirt
II
III
IV
-0,3 GTDC 90
BDC
270 ITDC 450 BDC Crank angle [◦ CA]
630
GTDC
Fig. 11: Piston secondary movement and tilt angle
therefore to regulate the lubrication supply. In Figure 8, the results of the floating liner measurement are plotted for the operation point of 2000 rpm and an indicated mean effective pressure of 6 bar. The upper plot illustrates the pressure curve of the floating liner engine and the second engine for LIF and piston-sided measurements. The difference between these curves during the combustion does not affect the general comparability of the experiments. The largest difference between the initial configuration and the configuration with reduced tangential force appears in the areas of the top and bottom dead center. The relative speed between the piston assembly and the cylinder liner is low, and the effect on the total friction loss is low. The oil film thickness and thereby the friction of the piston rings and the piston skirt is dependent on the available oil supply, which is regulated by the oil control ring. With reduced tangential force of the oil control ring, the available oil supply increases and the boundary friction is reduced. Compared to Figure 1, the friction force rises with higher relative speed as well as its impact on the total friction loss of the piston assembly. However, the difference between the two friction curves is reduced, which is plotted in Figure 8 as difference in [N]. In the following, the point of high relative speed is examined in detail for every stroke. The measured positions are visualized by vertical lines for the piston rings at nearly top speed (green vertical line) in LIF measuring point 1 and by the grey area where the piston skirt passes the LIF measuring point 2 (see also Figure
4. Results 4.1. Change of tangential force at the oil scraper ring Changing the tangential force of the oil control ring is a common experiment to analyze the trade-off effect between oil consumption and friction loss. The major task of the oil control ring is to regulate the oil supply between pistons rings and cylinder liner as well as between piston skirt and cylinder liner. The surface pressure between the oil control ring lands and the cylinder liner is high because of its low component height compared to the compression rings. This is essential in order to be able to follow the bore distortion of the cylinder liner, and 7
[bar]
30
Engine for LIF and piston-sided measurements
Floating liner engine
I
10
II
III
IV
p
C
20
Difference [N]
0 10 0 -10 -20 -30 OSR
PR1 vmax Piston
PR2
15-19%
Piston skirt
15-25%
28%
30%
22-28%
20-26%
12-22%
12-16%
100 Reference
80
Reduced FT
Friction force [N]
60 40 20 0 -20 -40 -60 -80 GTDC
90
BDC
270
ITDC
450
BDC
630
GTDC
Crank angle [◦ CA]
direction of movement
bot
top
top
bot
bot
top
top
50
bot
Fig. 8: Indicated cylinder pressure and friction-force
Reference Reduced FT
40 Film thickness [µm]
top ATS
30
TS
20
top
top
10
top bot
bot
bot
bot
0 70
80
90 I
100
260
270 II
280
290
430
Crank angle [◦ CA]
440
450 III
460
620
630
640 IV
Fig. 9: Oil film thickness measurements between piston skirt and cylinder liner, top/bot: top/bottom edge of the piston skirt
8
650
Film thickness [µm]
60 50 40 30 20 10
Film thickness [µm]
0
34 32 30 28 26 24 22 20 18 16 14 12 10 8 6 4 2 0
OSR
PR2
60
65
PR1
PR1
70
PR2
290
295
OSR
OSR
PR2
300 420 Crank angle [◦ CA]
PR1
425
PR1
430
650
OSR OSR
PR2
13%
PR2 15%
PR1
19%
PR1
Reference Reduced FT
OSR
OSR
17%
12% 27% 29%
17%
20%
18%
18%
13%
9% PR2
I
660
0.5%
8% 9%
OSR
PR2 655
II
PR1
PR1
III
PR2 IV
Fig. 10: Oil film thickness measurements between piston rings and cylinder liner
4). The shear stress ratio in the oil film decreases with a growing height of the oil film. With increasing speed the lubricant film gains a higher load bearing capacity, and the friction force difference of the configurations is reduced compared to the dead centers. Nevertheless, differences in lubricating oil film thickness are measured as shown in Figure 9 (LIF measuring point 2) and Figure 10 (LIF measuring point 1). The percentage reduction of the friction force is between 15 and 30 % (Figure 8). The decrease in lubricating film thickness between piston rings and cylinder liner varies between 9 to 29 % (Figure 10). The hydrodynamic friction reaches a higher level during the power stroke than in the other strokes for the examined points. The cylinder pressure acts on the compression rings and the piston. Consequently, the surface pressure between the piston rings and the cylinder liner is major, and the level of lubricant film thickness is reduced to 4 to 8 µm for the compression rings and 13 to 17 µm under the oil scraper ring (Figure 10), which is represented in stroke III of Figure 8. In Figure 9, the oil film thickness between the piston skirt and the liner is higher for the first two-thirds of the piston skirt in the power stroke (III) than in the intake stroke (I), whereas the oil film under the piston rings is lowest over the whole working cycle. This contrary effect can be explained by the piston secondary movement, see [33], pictured in Figure 11. The figure shows the position of three eddy current sensors which were installed on the
piston. Two sensors are placed on the thrust side – one in the ring land, one on the piston skirt – and a further sensor is mounted on the anti-thrust side. The calculation of the piston tilt angle is done by means of these three sensors, and it is exemplified by both the plot and the four pictures of the tilted pistons. While the piston skirt passes the LIF measuring point 2 (grey colored area, see also Figure 4), the measured distances on the piston skirt show slight distance or overlap. Only in the power stroke the thrust side of the piston skirt has a distinct distance to the cylinder liner. This leads to a tilt angle of -0.3◦ to the trust side according to the coordinate system in Figure 11 and a slight gap of the ring land to the liner. Between 70.5 and 104◦ CA and 430.5 and 464◦ CA, respectively, the piston skirt – moving downward – passes the measurement point 2. Figure 9 displays – in addition to the tilting position of the piston from Figure 11 – that the film thickness at the bottom edge of the piston is lower in the intake stroke (I) than in the power cycle (III) because the secondary movement of the piston leads to a slighter gap in this area at the beginning of the intake stroke (I). While the top edge moves past the measuring point, the piston top is less tilted to the thrust side in stroke I than in stroke III. Accordingly, the lubricating oil film thickness is now higher in stroke I than in stroke III. In upward direction (stroke II & IV) the oil film thickness is higher than in downward direction (stroke I & III) at the top edges of the piston because the piston top is tilted to 9
the anti-thrust side in contrast to the strokes I and III. As the bottom edge passes the measuring point, the film thickness is lower in stroke II and IV than in stroke III, wich can be explained by the bigger tilt angle (positive). Generally, the lubricating oil film reaches a maximum at the middle of the piston skirt, where the level is higher in upward direction just as the reduced tangential force of the oil scraper ring leads to a higher level over the entire measured area except for the area before and after the graphite-coated skirt (grey). There the signal level exceeds the measuring range maximum due to the smaller diameter of the piston (chamfers) resulting in a high amount of oil. Figure 10 illustrates the shift of the lubricating oil film thickness in consequence of the lower tangential force of the oil scraper ring among the piston rings and the cylinder liner. The lubricating oil film thickness is lower for all four strokes in the measuring point 1 (see also Figures 4 and 8) at high speed. Within the power and exhaust stroke (III and IV) the level of film thickness is generally lower than in the intake (I) and the compression stroke (II). This is justified by the high pressure level in the combustion chamber in stroke III that leads to a high surface pressure between the first piston ring and the cylinder liner, as well as between the second piston ring and the cylinder liner. This is explained by the higher friction force in stroke III, featured in Figure 8. By scraping a lot of oil from the liner in the power stroke, the level of oil film thickness is low until the end of the stroke IV when the piston cooling jet splashes fresh oil on the liner surface. Hence the amount of oil between the piston rings and the liner is higher in the intake and compression stroke. The area in front of the oil scraper ring is plotted in the upper graph of Figure 10. Moving downward (stroke I and III), the oil control ring scrapes too much oil from the liner, so that the signal level is too high to distinguish a difference between the oil film thicknesses in front of the oil scraper, regarding the different tangential forces. The area inside the piston rings is filled with oil. However there is no significant disparity of the signal level concerning the two different tangential forces. On the top land of the piston, the lubricating oil film thickness reaches a minimum inside the examined area in this figure. There is no significant observable difference for the measurement of the blow-by between the varied tangential forces of the oil scraper ring.
BDC
ITDC
BDC
TDC pcc p12,CF D−sim p23,CF D−sim p12,ind p23,ind
20 15 p [bar]
BDC
10 5 0 −180 −90
0
90
180
270
360
450
540
Crank angle [°CA]
Fig. 12: Comparison of simulated and indicated pressure curves acc. to [27]
measurements at the research engine, the ring motion of the 1D-simulation can be validated. The comparison of the ring land pressures between the rings can be carried out with the CFD results. In Figure 12 the measured ring land pressure is compared with the calculation results of the CFD simulation. Basically, there is a very good match in the trends, peaks and gradients recognizable. The first ring land pressure of the CFD simulation decreases a little slower after the maximum than the pressure curve from the measurement. This is a result of the neglection of the ring end gaps in the modeling of the CFD simulation (considering a 2D cut). The missing ring end gaps intensify the slower pressure decrease at small pressure ratios over the rings. A further possibility to validate the results is given by the blow-by. The integral blow-by at the research engine is determined by pressure differences over an orifice. The l measured result corresponds with a value of 3.42 min very l good with the simulation’s 3.45 min . In addition, the CFD simulation delivers a variety of locally resolved results. Exemplarily some selected flow results are presented in this work (fuel-lubricant interaction is not considered). The velocity field in the ring pack for selected crank angles is shown in Figure 13. The propagation of the flow through the ring pack during the high-pressure phase of an internal combustion engine can be seen very well in this figure. As a result of the high velocities at the ring flanks, a strong development of vortexes and turbulences behind the respective rings can be observed. A further flow path is determined between the rings and the cylinder liner. Because of surface roughness and the fact that the rings do not have the ability to perfectly fill the deformed liner, there are some small gaps between the cylinder liner and the rings. These small gaps and the high pressure differences can lead to high velocities between the rings and the liner.
4.2. CFD simulation compared to measurement As previously mentioned, the results of the simulation can be validated in different ways by the measurement data of the research engine. The ring-motion, that can be calculated in the 1D simulation, is a very important input for the CFD simulation. A good understanding of the ring motion provides the basis for a detailed calculation of the flow in the ring pack. Since the axial motion of the piston rings in the ring pack is available by 10
−30 °CA
0 °CA
30 °CA
270 °CA
60 °CA
v
330 °CA
390 °CA
450 °CA
αoil [−] 1.0 0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0
m s
20 18 16 14 12 10 8 6 4 2 0
Fig. 14: Oil distribution in the piston ring pack (with additional oil inlet) at selected crank angles acc. to [27]
Fig. 13: Velocity field of the gas flow simulation at selected crank angles acc. to [27]
lation method can achieve an important contribution to understand the processes in the ring pack.
Finally some selected results of the oil distribution in the ring pack are shown in Figure 14. Based on the initial distribution of the oil in the ring pack, the oil flow is mainly pushed by inertia forces and the interaction with the gas flow. The alteration of the oil in the oil reservoirs behind the rings can be seen very well in Figure 14, depending on the acceleration vector of the piston. Since where are only minor differences in pressure between combustion chamber and crankcase, the oil in the phase around GTDC is mainly influenced by inertia forces. Oil accumulations on the top land and behind the top ring respectively can easily reach the combustion chamber in this stage, which shows a source of higher oil consumption (cf. Figure 14). It is assumed that in the piston downward movement the area below the oil ring is fully filled with oil, which was considered in the simulation (Figure 14 at 390 and 450 ◦ CA). It is possible to prove single oil transport phenomena with this simulation method [27]. The knowledge of the oil distribution in the ring pack is very important for tribological calculations, but also for all appropriate 1D flow calculations. Besides the improvement of transport models in general (oil transport models, discharge coefficients, etc.), the prediction of oil throw-off at the top land can also be enhanced through the CFD simulation, and thereby input can be provided for simulation of the oil consumption. This new simu-
Taking all measured data into account, the different tribological conditions in consequence of the varied tangential forces can be represented and explained. For that reason, it is essential to improve the calculation of the oil transport and the blow-by gas in the piston ring pack for both comprehension and forecast of the tribological states of the piston assembly. This is crucial to optimize future engines in terms of fuel and oil consumption as well as acoustics and durability. 5. Conclusions An optical setup was developed for LIF measurement of the lubricating oil film thickness. A floating liner system was designed for crank angle resolved friction force measurement of the piston assembly. A production piston was modified to measure the piston secondary movement, the axial piston ring movement and the ring land pressure between the first and the second piston ring as well as between the second piston ring and the oil scraper ring.
11
Crank angle resolved friction force and lubricating oil film thickness could be measured and compared, including a variation of different tangential forces at the oil scraper ring. Effects on blow-by gas flow were not detected.
[7] P. Obert, T. M¨ uller, H.-J. F¨ ußer, and D. Bartel, “The influence of oil supply and cylinder liner temperature on friction, wear and scuffing behavior of piston ring cylinder liner contacts – a new model test,” Tribology International, vol. 94, pp. 306–314, 2016. [8] Goetze, Kolbenring-Handbuch. Federal-Mogul-Burscheid GmbH, 2003. [9] Stribeck R., “Die wesentlichen eigenschaften der gleit- und rollenlager,” Zeitschrift des VDI, no. Bd. 46, pp. 1341–1348, 1902. [10] U. Kessen, Tribologische Untersuchungen an der Kolbengruppe eines Dieselmotors bei hohen Mitteldr¨ ucken. PhD thesis, Universit¨ at Hannover, Hannover, 1999. [11] M. K. Werner, S. Graf, A. Merkle, and G. Wachtmeister, “Direct measurement of the piston assembly friction,” MTZ Worldwide, vol. 75, no. 1, pp. 50–57, 2014. [12] G. Garcia-Atance Fatjo, E. H. Smith, and I. Sherrington, “Mapping lubricating film thickness, film extent and ring twist for the compression-ring in a firing internal combustion engine,” Tribology International, vol. 70, pp. 112–118, 2014. [13] P. Saad, L. Kamo, M. Mekari, W. Bryzik, V. Wong, N. Dmitrichenko, and R. Mnatsakanov, “Modeling and measurement of tribological parameters between piston rings and liner in turbocharged diesel engine,” SAE, no. 2007-01-1440, 2007. [14] B. Littlefair, De la Cruz, M., S. Theodossiades, R. Mills, S. Howell-Smith, H. Rahnejat, and R. S. Dwyer-Joyce, “Transient tribo-dynamics of thermo-elastic compliant highperformance piston skirts,” Tribology Letters, vol. 53, no. 1, pp. 51–70, 2014. [15] R. Golloch, U. Kessen, and G. P. Merker, “Tribological investigations on the piston assembly group of a diesel engine,” MTZ Worldwide, vol. 63, no. 6, pp. 21–24, 2002. [16] T. Sato, H. Kurita, A. Ito, and H. Iwasaki, “Friction measurement of al-17%si monolithic with using newly developed floating liner device,” SAE Int J. Engines 8 (1): 2015, vol. 2015. [17] R. A. Mufti and M. Priest, “Experimental evaluation of pistonassembly friction under motored and fired conditions in a gasoline engine,” Journal of Tribology, vol. 127, no. 4, p. 826, 2005. [18] G. A. Livanos and N. P. Kyrtatos, “Friction model of a marine diesel engine piston assembly,” Tribology International, vol. 40, no. 10-12, pp. 1441–1453, 2007. [19] M. Werner, A. Merkle, S. Graf, R. Holzm¨ uller, and G. Wachtmeister, “Calculation of the piston assembly friction: Classification, validation and interpretation,” SAE, no. 2012-01-1323, 2012. [20] O. R. Ergen, G. Kurnaz, N. G. Soydemir, and O. Akalin, “Reduced oil consumption by laser surface texturing on cylinders,” SAE International Journal of Commercial Vehicles, vol. 1, no. 1, pp. 446–453, 2009. [21] D. Madden, K. Kim, and M. Takiguchi, “Part 1: Piston friction and noise study of three different piston architectures for an automotive gasoline engine,” SAE, no. 2006-01-0427, 2006. [22] S. Furuhama, M. Hiruma, and M. Tsuzita, “Piston ring motion and its influence on engine tribology,” SAE, no. 790860, 1979. [23] M. Puthiya Veettil and F. Shi, “Cfd analysis of oil/gas flow in piston ring-pack,” SAE, no. 2011-01-1406, 2011. [24] H. Shahmohamadi, R. Rahmani, H. Rahnejat, P. King, and C. Garner, eds., Cavitating Flow in Engine Piston RingCylinder Liner Conjunction, vol. IMECE2013-62395, 2013. [25] H. Shahmohamadi, M. Mohammadpour, R. Rahmani, H. Rahnejat, C. P. Garner, and S. Howell-Smith, “On the boundary conditions in multi-phase flow through the piston ring-cylinder liner conjunction,” Tribology International, vol. 90, pp. 164–174, 2015. [26] A. Oliva, S. Held, A. Herdt, and G. Wachtmeister, “Numerical simulation of the gas flow through the piston ring pack of an internal combustion engine,” SAE, no. 2015-01-1302, 2015. [27] A. Oliva and S. Held, “Numerical simulation and validation of the flow in the piston ring pack of an internal combustion engine,” Tribology International, 2016. [28] W. Hentschel, A. Grote, and O. Langer, “Measurement of wall
Differences in the strokes could be explained by the piston secondary movement, pressure conditions and direction of movement. It is therefore useful to attain as much measured data as possible. The gas and oil flow in the piston ring pack were calculated with a newly developed CFD-simulation method. The simulation was validated with the measured data from the research engine (ring motion, ring land pressures and blow-by). The simulation results can provide new input for tribological calculations and improve the insights of the gas and oil transport processes in the piston ring pack.
6. Acknowledgements ¨ The research project Kolbenring-Oltransport (IGF-no. 17553) was encouraged by the Federal Ministry for Economic Affairs and Energy on the orders of the German Bundestag with the help of the German Federation of Industrial Research Associations e. V. (AIF) and the Research Association for Combustion Engines e. V. (FVV). The authors thank for the allowance and for the support of the user committee led by the chairman Dr.-Ing. A. Robota (Federal Mogul Burscheid GmbH).
7. References [1] European Parliament and Council, “Regulation (ec) no 443/2009 of the european parliament and of the council of 23 april 2009 setting emission performance standards for new passenger cars as part of the community’s integrated approach to reduce co2 emissions from light-duty vehicles. regulation (ec) no 443/2009,” 2009. [2] European Parliament and Council, “Regulation (eu) no 333/2014 of the european parliament and of the council of 11 march 2014 amending regulation (ec) no 443/2009 to define the modalities for reaching the 2020 target to reduce co2 emissions from new passenger cars: Regulation (eu) no 333/2014,” 2014. [3] C. Pastorello and G. Mellios, “Monitoring co2 emissions from passenger cars and vans in 2013,” 2014. [4] C.-S. Ernst, C. Harter, I. Olschewski, and L. Eckstein, “Co2 emission reduction potential for passenger cars and light commercial vehicles post 2020: Management summary 123320,” 2014. [5] K. Holmberg, P. Andersson, and A. Erdemir, “Global energy consumption due to friction in passenger cars,” Tribology International, vol. 47, pp. 221–234, 2012. [6] S. C. Tung and M. L. McMillan, “Automotive tribology overview of current advances and challenges for the future,” Tribology International, vol. 37, no. 7, pp. 517–536, 2004.
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[31]
[32]
[33]
[34]
[35]
[36]
[37]
[38]
[39]
film thickness in the intake manifold of a standard production si engine by a spectroscopic technique,” SAE, no. 972832, 1997. K. Nakayama, T. Seki, M. Takiguchi, T. Someya, and S. Furuhama, “The effect of oil ring geometry on oil film thickness in the circumferential direction of the cylinder,” SAE, no. 982578, 1998. H.-J. Weimar, Entwicklung eines laser-optischen Messsystems ¨ zur kurbelwinkelaufgel¨ osten Bestimmung der Olfilmdicke zwischen Kolbenring und Zylinderwand in einem Ottomotor. PhD thesis, Universit¨ at Karlsruhe, Karlsruhe, 2002. S. Park and J. B. Ghandhi, “Fuel film temperature and thickness measurements on the piston crown of a direct-injection spark-ignition engine,” SAE, no. 2005-01-0649, 2005. C. Stein, M. Budde, S. Krause, S. Brandt, and F. Schlerege, “Schmier¨ olemission und Gemischbildung: Beeinflussung der Schmier¨ olemission durch die Gemischbildung im Brennraum von Verbrennungsmotoren: Vorhaben Nr. 933,” FVV Abschlussberichte, no. 901, 2010. C. Kirner, B. Uhlig, A. Behn, and Feindt M., “Kolbenring¨ ¨ Oltransport: Oltransport durch die Kolbenringe: Vorhaben Nr. 1124,” FVV Abschlussberichte, no. 1072, 2015. ¨ und Kraftstoffschichten S. Wigger, Charakterisierung von Olin der Kolbengruppe mittels laserinduzierter Fluoreszenz. PhD thesis, Universit¨ at Duisburg-Essen, Duisburg-Essen, 2014. J. E. Yun and S. S. Kim, “An improved approach on the instantaneous imep method for piston-ring assembly friction force measurement,” JSME International Journal, no. Series II, Vol. 35. No. 2, 1992. S. Furuhama and S. Sasaki, “New device for the measurement of piston frictional forces in small engines,” SAE, no. 831284, 1983. F. Koch, U. Geiger, and F.-G. Hermsen, PIFFO - Piston Friction Force Measurements During Engine Operation. Warrendale, PA: SAE International, 1996. F. Koch, F.-G. Hermsen, H. Marckwardt, and F.-G. Haubner, Friction Losses of Combustion Engines - Measurements, Analysis and Optimization. 1999. M. K. Werner, Entwicklung eines Motorpr¨ ufstands zur Untersuchung der Kolbengruppenreibung und deren Haupteinflussgr¨ oßen. PhD thesis, Technische Universit¨ at M¨ unchen, M¨ unchen, 2014.
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PII: DOI: Reference:
S0301-679X(16)00139-0 http://dx.doi.org/10.1016/j.triboint.2016.03.005 JTRI4111
To appear in: Tribiology International Received date: 22 December 2015 Revised date: 27 February 2016 Accepted date: 5 March 2016 Cite this article as: C. Kirner, J. Halbhuber, B. Uhlig, A. Oliva, S. Graf and G. Wachtmeister, Experimental and simulative research advances in the piston assembly of an internal combustion engine, Tribiology International, http://dx.doi.org/10.1016/j.triboint.2016.03.005 This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting galley proof before it is published in its final citable form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental and simulative research advances in the piston assembly of an internal combustion engine C. Kirnera , J. Halbhubera , B. Uhliga , A. Olivaa , S. Grafa , G. Wachtmeistera a Institute
of Internal Combustion Engines, Technical University of Munich
Abstract The piston assembly is the most complex tribological system within the internal combustion engine. In order to fully exploit its optimization potential a high level of system understanding is required. Therefore, in this paper experimental studies of two single cylinder engines are combinded with CFD-simulation of the piston ring pack. The introduced measurement and simulation techniques enable a holistic approach to the investigation of the tribological conditions of the piston assembly. The results – like the comparison of crank-angle resolved oil film thickness to the piston assembly friction measurement – illustrate the implications of design parameters of the piston assembly on the functional parameters friction, oil consumption, blow-by, wear and acoustics and thereby permit purposive system optimization. Keywords: Piston assembly, Lubricating oil film thickness measurement, Floating liner engine, CFD piston rings 1. Introduction In 2009 the EU member states have committed themselves to reduce the emission of polluting greenhouse gases, especially CO2 , by at least 20 % by 2020 compared to 1990. Therefore emissions of 95 g CO2 /km are the limit for new registrations of motorcars as of 2020 [1, 2]. Compared to the determined average value of 126.7 g CO2 /km by the EEA (European Environment Agency) [3] in 2013, the automotive industry is faced with tremendous economic and technical challenges, as a study of the IKA (Institute for Automotive Engineering, RWTH Aachen) shows [4]. Hence the reduction of CO2 emissions and the related fuel consumption is the key task in the development of modern combustion engines. A technical method to attain these target values, is to improve the mechanical efficiency. According to Holmberg [5] between 4.3 and 7.8 % of the fuel energy is dissipated by the friction between the piston assembly and the cylinder liner. Accordingly, the optimization of the piston assembly provides a good opportunity for the development of improved future engines. To exploit the potential for improvement, a detailed knowledge of the tribological processes is essential. Due to the constantly changing conditions like viscosity, velocity and pressure, the piston assembly is one of the most complex tribological systems inside the internal combustion engine [6]. In addition to the friction, the design of the piston assembly significantly influences the oil consumption and blow-by losses. Furthermore, wear and acoustic effects have to be taken into account, especially with the contact alteration of the piston. The major challenge in optimizing the piston assembly is Preprint submitted to Tribology International
thus to find measures that optimize the friction without influencing other tribological phenomena unfavourably. An evaluation of new design features with regard to effects on the friction-force is not sufficient for an effective improvement of the system piston assembly. In fact, the development of innovative components requires the holistic examination of all influences on this tribological system.
2. Theoretical basics In this paper, effects of a component variation – preload of the oil scraper ring – on the piston assembly friction, blow-by gas flow and the lubricating oil film thickness are investigated. The latter is crucial for friction, oil consumption and wear. So far little research has been done on dependencies between all these phenomena on the piston assembly. In many publications only individual effects, such as the study of friction and wear behavior at a tribology test rig [7] have been investigated. Interactions between the tribological behavior and other effects, such as changes in blow-by that occur during the combustion cycle, have not been analyzed yet. During the strokes the boundary conditions like viscosity, velocity and pressure change and the piston assembly passes states from boundary friction to hydrodynamic friction, as shown in Figure 1. The hydrodynamic friction represents a significant contribution to the friction losses, which is defined by the product of actual friction force and relative speed. The piston stops for a very short period of March 17, 2016
Nomenclature d h n p R t v
distance lubricating oil film thickness engine speed pressure surface roughness time velocity
α η κ µ
volume fraction viscosity piston tilt angle friction coefficient
Film thickness h Friction coefficient µ
time at the top and bottom dead center where the relative speed is zero. In these areas of boundary friction the oil film thickness is almost zero and wear occurs. The range of high relative speed in between the dead centers is of particular interest since the significant losses arise here [8].
A B
C
anti-thrust side bottom dead center crank angle gas exchange top dead center tangential force ignition top dead center indicated mean effective pressure top dead center thrust side
I II III IV
intake compression power exhaust
temperature, viscosity and pressure conditions at the piston rings are different. To understand the different friction states, the knowledge of lubrication oil film thickness is elementary, which – according to Figure 1 – increases with higher speed while viscosity and pressure are maintained. As a consequence, it is sensible to measure and compare friction-force and lubricating oil film thickness between piston and cylinder liner. In [12] the lubricating oil film thickness between the piston ring and the cylinder liner as well as the twist of the piston ring were considered in a modern 1.6 liter gazoline engine. By quantifying the difference in voltage drop between rings, oil film, and liner the research paper [13] presents an oil film thickness measurement method. Paper [14] provides measurement of oil film thickness by using noninvasive ultrasonic sensors on a motocross engine. In [15] crank angle resolved friction-forces were measured on an utility vehicle piston assembly. Simultaneously the lubricating film thickness between the cylinder liner and a piston ring was determined by means of inductive sensors. In [16] the floating-liner principle was used to analyse the running in effect of the cylinder liner. In [17] a piston assembly friction measurement was carried out on a singlecylinder gasoline engine using the IMEP method. The calculation of the friction-forces – as described in [18, 19] – is possible with the knowledge of the crank angle resolved lubricating film thickness between the friction partners. The oil consumption plays an important role in the design of modern internal combustion engines, and it is a subject of current research. In [20] the oil consumption was detected by means of a sulphur tracer and weighing. Moreover the influence of different laser-textures on the cylinder liner was measured. In particular, the reverse blow-by contributes to the oil consumption, which in turn is influenced by ring land pressures and piston ring movement. So in addition to blow-by, piston ring movement and ring land pressure were measured to understand their
A: Boundary friction (h → 0) B: Mixed friction (h ≈ R) C: Hydrodynamic lubrication (h R)
0
ATS BDC CA GTDC FT ITDC IMEP TDC TS
Boundary friction Hydrodynamic friction 0 viscosity η · velocity v pressure p
Fig. 1: Film thickness and Stribeck curve acc. to [8, 9, 10]
The exemplary friction-force plot in Figure 2 illustrates that a separate consideration of the power cycle or other parts of the stroke is not effective in terms of an improvement concerning the tribological conditions. Thus, in order to understand the occurring phenomena on the piston assembly sufficiently, it is important that the complex effects are considered for the entire working cycle. According to [11] the dead centers indicate mixed friction conditions having different characteristics in TDC and BTDC, as liner 2
small influence of the relative ring end gap positions of the three piston rings to each other on the blow-by is detected. However, the position of the ring end gaps to each other has a greater influence on the oil transport. Other investigations are limited to the CFD-simulation of very local phenomena in the ring pack. The relative velocities between rings and cylinder liner during a working cycle and the low film height can cause cavitation effects on the ring running surface. These effects modify the tribological conditions between ring and cylinder liner and can have an impact on all development objectives (blowby, friction, oil transport and wear) in the piston ring pack [24, 25]. New studies focus on the calculation of the flow in the ring pack by means of a 2D-CFD approach where the ring movement is considered. This method has been successfully applied to a diesel and a gasoline engine [26, 27]. The calculation method, the achieved results and also the possibilities of validation of the simulation will be described in the following.
Friction force [N]
Boundary friction Hydrodynamic friction Overall friction
0
90
180
270 360 450 540 Crankangle [◦ CA]
630
720
Fig. 2: Schematic friction force acc. to [19, 11]
3. Method effect on oil consumption. The reduction of friction and blow-by amount is often a conflict of objectives which is the reason for measuring these parameters. In general, decreasing piston clearance results in better acoustics, but higher friction losses. Studies of correlations between friction- and acoustics-behavior were carried out in [21]. However, changes in lubricating oil film thickness and blow-by gas flow were not examined here. Because of its complex modelling, the field of CFD of the ring pack is still relatively young. The 0D- and 1D- simulations that are based on the chamber orifice principle have been established as state of the technology for a long time [22]. The advantages of this method are the simple modeling and short calculation times. However there are also a number of disadvantages. In order to achieve reliable results, the flow coefficients for the small gaps at the piston rings (cylinder liner, ring flanks and ring end gaps) must be specified or rather have to agree with the measured data for the blow-by. Also simplifications for oil transport in the ring pack are made. The mainly used Reynolds equations for thin films are not a valid assumption for calculating the oil transport in the ring pack in the areas behind or between the rings. The CFD simulation provides new knowledge to improve the simplified models in the 0D-/1D-simulation. A 3D CFD-simulation of the gas and oil flow in the ring pack has already been carried out in 2011, see [23]. In this study, the piston rings are considered rigid at certain positions of the ring groove. Depending on the position of the rings, the results of the blow-by and oil transport were significantly different. The ring motion could be identified as an important influencing factor for the transport phenomena in the ring pack, hence the piston ring motion should not be neglected. As a further result, only a
Measurements are carried out using two single-cylinder gasoline engines. Both engines have the same crank drive and cylinder capacity, as the engine data show in Table 1. Piston, piston rings and conrod are series production parts of a gasoline engine with a cylinder capacity of 2.0 liter. The cylinder liner design is close to the production engine. This design ensures that all results are compatible between the research engines and transferable to production engines.
Engine parameter
Value
Cylinder capacity
0.5 L
Bore
82.5 mm
Stroke
92.8 mm
Compression ratio
9.6:1
Operating points
Value
Engine speed
1500 rpm
IMEP
2000 rpm
5 bar
6 bar ◦
Oil temperature
90 C
Coolant temperature
90 ◦C
Table 1: Engine data
To prove the relation between piston assembly friction and oil film thickness, the tangential force of oil control ring is reduced (sec 4.1). This change of design makes it possible to consider the trade-off between friction reduction and oil consumption. The friction measurement is realized by a 3
floating liner engine (built up in 2010), and the measurement of the oil film thickness, ring land pressures, piston secondary, and piston ring movement is applied in a second engine (built up in 2014) at an engine speed of 2000 rpm and an IMEP of 6 bar, respectively (sec 3.1–3.3 and 4.1). The results of the CFD simulation (sec 3.4 and 4.2) are compared to the measurements at an engine speed of 1500 rpm and an IMEP of 5 bar, Table 1.
Probe Cylinder housing
O-ring
Liner Ferrule
Nut 1
3.1. LIF Measurement of lubricating oil film thickness Oil film thicknesses are measured using laser-induced fluorescence (LIF), similar to [28, 29, 30, 31, 32]. A monochromatic laser-light excites fluorescence in the lubricant and the added dye. Optical fibers are used for the transport of the laser- and fluorescent light. 2 a
b Snap ring
d
%
0 450
A prior calibration that allows to put oil film thicknesses in relation to the fluorescent intensity enables absolute measures of lubricant film thicknesses with the help of LIF. Calibration techniques that allow adjustment of lubricant film thicknesses and measurement of fluorescent intensity are shown in [30, 32, 34]. To take the adjustment of the optical setup and the quality of the measurement point on the liner (regarding polishing and assembly-offset) into account, a reference experiment was conducted directly on the experimental engine. Piston rings and precision foils with 12 µm thickness were used. An increasing number of foils (up to 4 with a common thickness of 48 µm) were clamped between the piston ring and the liner, and under the addition of fluorescent oil, a linear relation between the number of sheets and fluorescent intensity could be measured. The setup ensures that reflections on the piston rings are considered within the calibration procedure and the measurement. Furthermore, the influence of oil temperature, like in [30, 31] was investigated. By increasing the oil temperature from 20 ◦ C to 70 ◦ C, the fluorescence intensity decreases to 85 % of the value at 20 ◦ C. With a further raise to 90 ◦ C oil temperature, the fluorescence intensity was reduced to 75 % of the initial value at 20 ◦ C. The oil- and coolant temperature were set to 90 ◦ C in the testing scenario, for this reason the calibration figures for this temperature were applied in the signal processing. In order to measure the lubricating oil film thickness between the piston assembly and the cylinder liner in hydrodynamic lubrication state, two measuring points were placed on the liner, as displayed in Figure 4. Measuring
Exciting Laser Dye Absorbtion Dye Emission Dichroic Filter Emission Filter
50
500 Wavelength [nm]
Piston assembly
e
g f 100
Coolant
Fig. 4: LIF measuring points 1 and 2, piston assembly in vmax acc. to [33]
c c
Optical fiber
550
Fig. 3: Optical setup acc. to [33]
A continuous-wave laser with a wavelength of 473 nm leads to a sufficient excitation, and the emission spectrum is located at higher wavelengths due to the stokes-shift, with a maximum around 500 nm. Figure 3 shows one beam path of the optical setup for sixteen synchronous points of measurement. Laser light from a single laser (a) is divided by a cascade of beam-splitters into sixteen optical paths (b), each of which is coupled by two mirrors (c), a dichroic filter (d), and a fiber coupling (e) into a single fiber leading to the measurement point in the engine. There, laser light exits the fiber directly into the oil film and induces fluorescence that returns through the same fiber. Dichroic filters (d) and additional emission filters (f) separate fluorescent wavelengths from reflections. The fluorescent light is focused on a photodiode (g) and its intensity is measured. 4
point 1 is used for the examination of the lubricating oil film thickness between the piston rings and the cylinder liner at the hydrodynamic lubrication state. The piston rings have a speed of 96–99 % of top speed. Measuring point 2 is positioned in such a way that the illustrated position on the piston skirt is measured at the top speed of the piston. Probes are incorporated inside the cylinder housing and they are fixed with nuts to the cylinder liner. The waterproofing is realized by two radial caulking o-rings for one measuring point; one of them is for sealing between cylinder housing and probe and the other one for in between cylinder liner and probe. Advantages of the radial sealing between the cylinder liner and the probe include close-toproduction liner shapes and surfaces, as well as greatly reduced influence on the liner distortion with proper cooling. The optical fibers are glued in stainless steel ferrules and polished for both a high emission of the laser light and the coupling of the fluorescence light back to the fiber. To fit the ferrules flush with the liner surface, an aluminium cylinder is positioned inside the liner with a clearance fit and heated such that the clearance fit changes into a interference fit. This ensures that the ferrules are glued flush with the liner. Since the ferrules involving the fibers are only glued inside the probes, the dismounting and service of components in this assembly is possible without destroying the optical fibers.
6 FM
3
5
7 2 FF
8
FS 4 1: 2: 3: 4: 5: 6: 7: 8:
Cylinder liner Cooling jacket Load cell Radial bearing Basic engine Cylinder head Radial sealing Piston assembly
1
Fig. 5: Schematic floating liner diagram
is constant, operation point independent and compensated within the measurement data processing. Thereby the friction of the piston assembly can be determined isolatedly. The central design element of a floating liner engine is the sealing between the floating cylinder liner and the cylinder head. The combustion chamber has to be sealed without blocking the axial degree of freedom of the cylinder liner. In this case the cylinder head is extended into the cooling jacket, which has the same bore diameter as the cylinder liner. The radial sealant is realized by a high temperature gasket. With this concept, a gas force compensation, that is required for the mentioned Musashi and PIFFO concept is dispensable [22, 36, 37, 38]. For a precise measuring of the piston assembly friction force it is necessary to control disturbance variables like cylinder pressure and sealing, engine vibrations, crankcase pressure, and temperature deviations. With this test bench, the external conditioning and the employed measuring procedure, it is possible to handle these disturbance variables, and a high level of reproducibility is achieved. The reproducibility is described by the deviation of two measurements between which a simulated change of components is performed. Therefore the engine is disassembled after the first measuring cycle, as it is carried out for a change of components, but assembled with exactly the same components. After conditioning the engine the identical measurement cycle is operated, and the results are compared. To gain results of high reproducibility, a stationary operating point and equal center of combus-
3.2. Crank angle resolved measurement of piston assembly friction To measure the piston assembly friction in an internal combustion engine, different methods like strip method, indication method, IMEP method [35] and the floatingliner-method are available [36, 22]. Solely the floating liner method enables direct crank angle resolved measurements of the piston assembly friction under actual internal combustion engine conditions. The IMEP method, carried out in [17], allows the indirect crank angle resolved measurement of the piston assembly friction. The single cylinder test station is based on the same series engine as the LIF equipped engine whose basic data is summarized in Table 1. There are several design concepts for floating liner systems like the Musashi engine or the PIFFO engine, which are explained in detail at [16, 22, 36, 37, 38]. The basic concept of the second generation floating liner engine is pictured in Figure 5. The assembly of cylinder liner and cooling jacket is mounted on the basic engine by two load cells, which are positioned in the crank shaft axis. This arrangement of the load cells reduces the influence of lateral piston forces on the measurement results. The load cells are prestressed to allow the measurement of compressive and tractive forces. The lateral piston force is retained by a radial bearing. Hence the measuring device is mounted in shunt mode and a calibration is necessary to define the ratio between the applied friction force and the force measured by the load cells. The shunt force ratio 5
tion mass are required. The combustion cycle averaged standard deviation of the measured friction force curves is less than 2.5 N for every operating point of the performance map [11, 39]. This measurement device allows to examine variations of piston assembly components as well as changes of engine boundary conditions and to evaluate the results. 3.3. Setup for axial piston ring movement, piston secondary movement and ring land pressure measurement The measurement of the axial piston ring movement is conducted with the help of eddy current sensors. The position of the sensors is illustrated on the left-hand side of the Figure 6. For the integration of the sensors in the top land two through-holes, through which the sensors can be assembled, are manufactured. Both of them are sited in the ring land between the first and the second compression ring that leads to lower mechanical and thermal stresses for the sensor measuring the motion of the first piston ring. The sensor cables are drawn through radial bores into the piston. Both this radial bores and the through-holes in the top land are pasted up after the mounting. The sensor cables are routed via a mechanical linkage to the measurement data acquisition. On the right-hand side of Figure 6, the position of the resistive pressure sensors for the ring land pressure measurement and two eddy current sensors for recording the piston secondary movement are marked out. The pressure sensors have a diameter of 1.9 mm, and therefore a flush mounting to the outer surface is only possible on the first ring land. In the second ring land, however, no flush mounting is possible, which is why the installed sensors are recessed. One additional eddy current sensor for the detection of the piston secondary movement is installed on the anti-thrust side of the piston, see Figure 11.
Fig. 7: Simulation mesh and computational domain [27]
3.4. CFD-computation of the flow in the piston ring pack Basically, many preliminary considerations are necessary for the numerical calculation of the flow in the ring pack. The computational domain can be defined in different ways depending on the modeling depth. The known options are the 3D full model, a 3D segment and a 2D cut through the ring pack, [27]. The advantages of the 2D cut are the lower number of cells that are necessary for the description of the flow region and the simpler implementation of the ring motion (axial, radial, twist) that is important for the flow. For 3D simulations the complexity of the ring motion increases with a growing number of degrees of freedom. The disadvantage of the 2D simulation is the neglection of the ring end gaps. Despite this simplification, the flow phenomena can be represented well, [26, 27]. Before a computational mesh can be built, the relevant area for the flow has to be derived from CAD data. In addition to the deformation of the cylinder liner resulting from the assembly process, the thermal deformation of the components need to be considered and averaged for the 2D cut. A differentiation between the thrust and the anti-thrust side has not been made. The derived geometry and a computational mesh for the CFD-simulation can be seen in Figure 7. Through the narrow gaps on the ring flanks or rather between the rings and the cylinder liner, locally a fine res-
Eddy current sensors Through-hole
Sensor cables
Resistive pressure sensors
Eddy current sensors
Fig. 6: Axial ring movement (left), piston secondary movement and ring land pressure measurement setup (right) acc. to [33]
6
olution for the mesh has to be chosen. It was ensured that the computational mesh is widely orthogonal structured, which has numerical advantages as well as advantages for the mesh movement. A detailed description of the mesh movement can be found in [26]. The consideration of the ring motion in the CFD simulation guarantees a realistic depiction of the occurring effects. Pressure and temperature of the combustion camber on the one side and pressure in the crankcase on the other side are defined as boundary conditions of the simulation. Moreover, the temperatures at the cylinder liner, the piston and the rings (averaged over one working cycle) are specified. The temperature data may be derived from FEM-simulations or, as in this case, from measurements on the research engine. The motion of the rings is externally calculated in a 1D simulation and then set in the CFD simulation. The movement of the piston is taken into account by a transformation of the coordinate system. The cylinder liner moves from the perspective of the piston-fixed coordinate system with negative piston velocity. The CFD simulation also includes the oil in the piston ring pack. For this exchange models have to be defined between the phases air and oil. The distribution of the oil in the ring pack can be calculated by the initialization of oil in the reservoirs behind the rings. In order to avoid the ring pack running dry the simulation assumes that in the piston downward movement the area below the oil ring is fully filled with oil. The aim of this multi-phase simulation is to calculate the oil distribution in the ring pack or rather the calculation of oil transport phenomena, which has been successfully demonstrated in [27]. The validation of the flow results can be carried out in various ways by the available measurement results from the research engine. The measured ring motion at the engine can be used to validate the ring motion results of the 1D simulation. The ring land pressures represent another important validation opportunity to calculate the correct pressure ratios over the rings. Finally, the integral result of the blow-by from the CFD-simulation can also be validated with the measured blow-by volume flow rate.
κ +
+
-
dT S
piston ring land
dT S
piston skirt
dAT S
coordinate system
piston skirt
κ
ATS
TS
0,3
dT S
piston ring land
d [mm], κ [°]
0,2 0,1
dT S
piston skirt
0 -0,1 -0,2
dAT S
I
κ piston skirt
II
III
IV
-0,3 GTDC 90
BDC
270 ITDC 450 BDC Crank angle [◦ CA]
630
GTDC
Fig. 11: Piston secondary movement and tilt angle
therefore to regulate the lubrication supply. In Figure 8, the results of the floating liner measurement are plotted for the operation point of 2000 rpm and an indicated mean effective pressure of 6 bar. The upper plot illustrates the pressure curve of the floating liner engine and the second engine for LIF and piston-sided measurements. The difference between these curves during the combustion does not affect the general comparability of the experiments. The largest difference between the initial configuration and the configuration with reduced tangential force appears in the areas of the top and bottom dead center. The relative speed between the piston assembly and the cylinder liner is low, and the effect on the total friction loss is low. The oil film thickness and thereby the friction of the piston rings and the piston skirt is dependent on the available oil supply, which is regulated by the oil control ring. With reduced tangential force of the oil control ring, the available oil supply increases and the boundary friction is reduced. Compared to Figure 1, the friction force rises with higher relative speed as well as its impact on the total friction loss of the piston assembly. However, the difference between the two friction curves is reduced, which is plotted in Figure 8 as difference in [N]. In the following, the point of high relative speed is examined in detail for every stroke. The measured positions are visualized by vertical lines for the piston rings at nearly top speed (green vertical line) in LIF measuring point 1 and by the grey area where the piston skirt passes the LIF measuring point 2 (see also Figure
4. Results 4.1. Change of tangential force at the oil scraper ring Changing the tangential force of the oil control ring is a common experiment to analyze the trade-off effect between oil consumption and friction loss. The major task of the oil control ring is to regulate the oil supply between pistons rings and cylinder liner as well as between piston skirt and cylinder liner. The surface pressure between the oil control ring lands and the cylinder liner is high because of its low component height compared to the compression rings. This is essential in order to be able to follow the bore distortion of the cylinder liner, and 7
[bar]
30
Engine for LIF and piston-sided measurements
Floating liner engine
I
10
II
III
IV
p
C
20
Difference [N]
0 10 0 -10 -20 -30 OSR
PR1 vmax Piston
PR2
15-19%
Piston skirt
15-25%
28%
30%
22-28%
20-26%
12-22%
12-16%
100 Reference
80
Reduced FT
Friction force [N]
60 40 20 0 -20 -40 -60 -80 GTDC
90
BDC
270
ITDC
450
BDC
630
GTDC
Crank angle [◦ CA]
direction of movement
bot
top
top
bot
bot
top
top
50
bot
Fig. 8: Indicated cylinder pressure and friction-force
Reference Reduced FT
40 Film thickness [µm]
top ATS
30
TS
20
top
top
10
top bot
bot
bot
bot
0 70
80
90 I
100
260
270 II
280
290
430
Crank angle [◦ CA]
440
450 III
460
620
630
640 IV
Fig. 9: Oil film thickness measurements between piston skirt and cylinder liner, top/bot: top/bottom edge of the piston skirt
8
650
Film thickness [µm]
60 50 40 30 20 10
Film thickness [µm]
0
34 32 30 28 26 24 22 20 18 16 14 12 10 8 6 4 2 0
OSR
PR2
60
65
PR1
PR1
70
PR2
290
295
OSR
OSR
PR2
300 420 Crank angle [◦ CA]
PR1
425
PR1
430
650
OSR OSR
PR2
13%
PR2 15%
PR1
19%
PR1
Reference Reduced FT
OSR
OSR
17%
12% 27% 29%
17%
20%
18%
18%
13%
9% PR2
I
660
0.5%
8% 9%
OSR
PR2 655
II
PR1
PR1
III
PR2 IV
Fig. 10: Oil film thickness measurements between piston rings and cylinder liner
4). The shear stress ratio in the oil film decreases with a growing height of the oil film. With increasing speed the lubricant film gains a higher load bearing capacity, and the friction force difference of the configurations is reduced compared to the dead centers. Nevertheless, differences in lubricating oil film thickness are measured as shown in Figure 9 (LIF measuring point 2) and Figure 10 (LIF measuring point 1). The percentage reduction of the friction force is between 15 and 30 % (Figure 8). The decrease in lubricating film thickness between piston rings and cylinder liner varies between 9 to 29 % (Figure 10). The hydrodynamic friction reaches a higher level during the power stroke than in the other strokes for the examined points. The cylinder pressure acts on the compression rings and the piston. Consequently, the surface pressure between the piston rings and the cylinder liner is major, and the level of lubricant film thickness is reduced to 4 to 8 µm for the compression rings and 13 to 17 µm under the oil scraper ring (Figure 10), which is represented in stroke III of Figure 8. In Figure 9, the oil film thickness between the piston skirt and the liner is higher for the first two-thirds of the piston skirt in the power stroke (III) than in the intake stroke (I), whereas the oil film under the piston rings is lowest over the whole working cycle. This contrary effect can be explained by the piston secondary movement, see [33], pictured in Figure 11. The figure shows the position of three eddy current sensors which were installed on the
piston. Two sensors are placed on the thrust side – one in the ring land, one on the piston skirt – and a further sensor is mounted on the anti-thrust side. The calculation of the piston tilt angle is done by means of these three sensors, and it is exemplified by both the plot and the four pictures of the tilted pistons. While the piston skirt passes the LIF measuring point 2 (grey colored area, see also Figure 4), the measured distances on the piston skirt show slight distance or overlap. Only in the power stroke the thrust side of the piston skirt has a distinct distance to the cylinder liner. This leads to a tilt angle of -0.3◦ to the trust side according to the coordinate system in Figure 11 and a slight gap of the ring land to the liner. Between 70.5 and 104◦ CA and 430.5 and 464◦ CA, respectively, the piston skirt – moving downward – passes the measurement point 2. Figure 9 displays – in addition to the tilting position of the piston from Figure 11 – that the film thickness at the bottom edge of the piston is lower in the intake stroke (I) than in the power cycle (III) because the secondary movement of the piston leads to a slighter gap in this area at the beginning of the intake stroke (I). While the top edge moves past the measuring point, the piston top is less tilted to the thrust side in stroke I than in stroke III. Accordingly, the lubricating oil film thickness is now higher in stroke I than in stroke III. In upward direction (stroke II & IV) the oil film thickness is higher than in downward direction (stroke I & III) at the top edges of the piston because the piston top is tilted to 9
the anti-thrust side in contrast to the strokes I and III. As the bottom edge passes the measuring point, the film thickness is lower in stroke II and IV than in stroke III, wich can be explained by the bigger tilt angle (positive). Generally, the lubricating oil film reaches a maximum at the middle of the piston skirt, where the level is higher in upward direction just as the reduced tangential force of the oil scraper ring leads to a higher level over the entire measured area except for the area before and after the graphite-coated skirt (grey). There the signal level exceeds the measuring range maximum due to the smaller diameter of the piston (chamfers) resulting in a high amount of oil. Figure 10 illustrates the shift of the lubricating oil film thickness in consequence of the lower tangential force of the oil scraper ring among the piston rings and the cylinder liner. The lubricating oil film thickness is lower for all four strokes in the measuring point 1 (see also Figures 4 and 8) at high speed. Within the power and exhaust stroke (III and IV) the level of film thickness is generally lower than in the intake (I) and the compression stroke (II). This is justified by the high pressure level in the combustion chamber in stroke III that leads to a high surface pressure between the first piston ring and the cylinder liner, as well as between the second piston ring and the cylinder liner. This is explained by the higher friction force in stroke III, featured in Figure 8. By scraping a lot of oil from the liner in the power stroke, the level of oil film thickness is low until the end of the stroke IV when the piston cooling jet splashes fresh oil on the liner surface. Hence the amount of oil between the piston rings and the liner is higher in the intake and compression stroke. The area in front of the oil scraper ring is plotted in the upper graph of Figure 10. Moving downward (stroke I and III), the oil control ring scrapes too much oil from the liner, so that the signal level is too high to distinguish a difference between the oil film thicknesses in front of the oil scraper, regarding the different tangential forces. The area inside the piston rings is filled with oil. However there is no significant disparity of the signal level concerning the two different tangential forces. On the top land of the piston, the lubricating oil film thickness reaches a minimum inside the examined area in this figure. There is no significant observable difference for the measurement of the blow-by between the varied tangential forces of the oil scraper ring.
BDC
ITDC
BDC
TDC pcc p12,CF D−sim p23,CF D−sim p12,ind p23,ind
20 15 p [bar]
BDC
10 5 0 −180 −90
0
90
180
270
360
450
540
Crank angle [°CA]
Fig. 12: Comparison of simulated and indicated pressure curves acc. to [27]
measurements at the research engine, the ring motion of the 1D-simulation can be validated. The comparison of the ring land pressures between the rings can be carried out with the CFD results. In Figure 12 the measured ring land pressure is compared with the calculation results of the CFD simulation. Basically, there is a very good match in the trends, peaks and gradients recognizable. The first ring land pressure of the CFD simulation decreases a little slower after the maximum than the pressure curve from the measurement. This is a result of the neglection of the ring end gaps in the modeling of the CFD simulation (considering a 2D cut). The missing ring end gaps intensify the slower pressure decrease at small pressure ratios over the rings. A further possibility to validate the results is given by the blow-by. The integral blow-by at the research engine is determined by pressure differences over an orifice. The l measured result corresponds with a value of 3.42 min very l good with the simulation’s 3.45 min . In addition, the CFD simulation delivers a variety of locally resolved results. Exemplarily some selected flow results are presented in this work (fuel-lubricant interaction is not considered). The velocity field in the ring pack for selected crank angles is shown in Figure 13. The propagation of the flow through the ring pack during the high-pressure phase of an internal combustion engine can be seen very well in this figure. As a result of the high velocities at the ring flanks, a strong development of vortexes and turbulences behind the respective rings can be observed. A further flow path is determined between the rings and the cylinder liner. Because of surface roughness and the fact that the rings do not have the ability to perfectly fill the deformed liner, there are some small gaps between the cylinder liner and the rings. These small gaps and the high pressure differences can lead to high velocities between the rings and the liner.
4.2. CFD simulation compared to measurement As previously mentioned, the results of the simulation can be validated in different ways by the measurement data of the research engine. The ring-motion, that can be calculated in the 1D simulation, is a very important input for the CFD simulation. A good understanding of the ring motion provides the basis for a detailed calculation of the flow in the ring pack. Since the axial motion of the piston rings in the ring pack is available by 10
−30 °CA
0 °CA
30 °CA
270 °CA
60 °CA
v
330 °CA
390 °CA
450 °CA
αoil [−] 1.0 0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0
m s
20 18 16 14 12 10 8 6 4 2 0
Fig. 14: Oil distribution in the piston ring pack (with additional oil inlet) at selected crank angles acc. to [27]
Fig. 13: Velocity field of the gas flow simulation at selected crank angles acc. to [27]
lation method can achieve an important contribution to understand the processes in the ring pack.
Finally some selected results of the oil distribution in the ring pack are shown in Figure 14. Based on the initial distribution of the oil in the ring pack, the oil flow is mainly pushed by inertia forces and the interaction with the gas flow. The alteration of the oil in the oil reservoirs behind the rings can be seen very well in Figure 14, depending on the acceleration vector of the piston. Since where are only minor differences in pressure between combustion chamber and crankcase, the oil in the phase around GTDC is mainly influenced by inertia forces. Oil accumulations on the top land and behind the top ring respectively can easily reach the combustion chamber in this stage, which shows a source of higher oil consumption (cf. Figure 14). It is assumed that in the piston downward movement the area below the oil ring is fully filled with oil, which was considered in the simulation (Figure 14 at 390 and 450 ◦ CA). It is possible to prove single oil transport phenomena with this simulation method [27]. The knowledge of the oil distribution in the ring pack is very important for tribological calculations, but also for all appropriate 1D flow calculations. Besides the improvement of transport models in general (oil transport models, discharge coefficients, etc.), the prediction of oil throw-off at the top land can also be enhanced through the CFD simulation, and thereby input can be provided for simulation of the oil consumption. This new simu-
Taking all measured data into account, the different tribological conditions in consequence of the varied tangential forces can be represented and explained. For that reason, it is essential to improve the calculation of the oil transport and the blow-by gas in the piston ring pack for both comprehension and forecast of the tribological states of the piston assembly. This is crucial to optimize future engines in terms of fuel and oil consumption as well as acoustics and durability. 5. Conclusions An optical setup was developed for LIF measurement of the lubricating oil film thickness. A floating liner system was designed for crank angle resolved friction force measurement of the piston assembly. A production piston was modified to measure the piston secondary movement, the axial piston ring movement and the ring land pressure between the first and the second piston ring as well as between the second piston ring and the oil scraper ring.
11
Crank angle resolved friction force and lubricating oil film thickness could be measured and compared, including a variation of different tangential forces at the oil scraper ring. Effects on blow-by gas flow were not detected.
[7] P. Obert, T. M¨ uller, H.-J. F¨ ußer, and D. Bartel, “The influence of oil supply and cylinder liner temperature on friction, wear and scuffing behavior of piston ring cylinder liner contacts – a new model test,” Tribology International, vol. 94, pp. 306–314, 2016. [8] Goetze, Kolbenring-Handbuch. Federal-Mogul-Burscheid GmbH, 2003. [9] Stribeck R., “Die wesentlichen eigenschaften der gleit- und rollenlager,” Zeitschrift des VDI, no. Bd. 46, pp. 1341–1348, 1902. [10] U. Kessen, Tribologische Untersuchungen an der Kolbengruppe eines Dieselmotors bei hohen Mitteldr¨ ucken. PhD thesis, Universit¨ at Hannover, Hannover, 1999. [11] M. K. Werner, S. Graf, A. Merkle, and G. Wachtmeister, “Direct measurement of the piston assembly friction,” MTZ Worldwide, vol. 75, no. 1, pp. 50–57, 2014. [12] G. Garcia-Atance Fatjo, E. H. Smith, and I. Sherrington, “Mapping lubricating film thickness, film extent and ring twist for the compression-ring in a firing internal combustion engine,” Tribology International, vol. 70, pp. 112–118, 2014. [13] P. Saad, L. Kamo, M. Mekari, W. Bryzik, V. Wong, N. Dmitrichenko, and R. Mnatsakanov, “Modeling and measurement of tribological parameters between piston rings and liner in turbocharged diesel engine,” SAE, no. 2007-01-1440, 2007. [14] B. Littlefair, De la Cruz, M., S. Theodossiades, R. Mills, S. Howell-Smith, H. Rahnejat, and R. S. Dwyer-Joyce, “Transient tribo-dynamics of thermo-elastic compliant highperformance piston skirts,” Tribology Letters, vol. 53, no. 1, pp. 51–70, 2014. [15] R. Golloch, U. Kessen, and G. P. Merker, “Tribological investigations on the piston assembly group of a diesel engine,” MTZ Worldwide, vol. 63, no. 6, pp. 21–24, 2002. [16] T. Sato, H. Kurita, A. Ito, and H. Iwasaki, “Friction measurement of al-17%si monolithic with using newly developed floating liner device,” SAE Int J. Engines 8 (1): 2015, vol. 2015. [17] R. A. Mufti and M. Priest, “Experimental evaluation of pistonassembly friction under motored and fired conditions in a gasoline engine,” Journal of Tribology, vol. 127, no. 4, p. 826, 2005. [18] G. A. Livanos and N. P. Kyrtatos, “Friction model of a marine diesel engine piston assembly,” Tribology International, vol. 40, no. 10-12, pp. 1441–1453, 2007. [19] M. Werner, A. Merkle, S. Graf, R. Holzm¨ uller, and G. Wachtmeister, “Calculation of the piston assembly friction: Classification, validation and interpretation,” SAE, no. 2012-01-1323, 2012. [20] O. R. Ergen, G. Kurnaz, N. G. Soydemir, and O. Akalin, “Reduced oil consumption by laser surface texturing on cylinders,” SAE International Journal of Commercial Vehicles, vol. 1, no. 1, pp. 446–453, 2009. [21] D. Madden, K. Kim, and M. Takiguchi, “Part 1: Piston friction and noise study of three different piston architectures for an automotive gasoline engine,” SAE, no. 2006-01-0427, 2006. [22] S. Furuhama, M. Hiruma, and M. Tsuzita, “Piston ring motion and its influence on engine tribology,” SAE, no. 790860, 1979. [23] M. Puthiya Veettil and F. Shi, “Cfd analysis of oil/gas flow in piston ring-pack,” SAE, no. 2011-01-1406, 2011. [24] H. Shahmohamadi, R. Rahmani, H. Rahnejat, P. King, and C. Garner, eds., Cavitating Flow in Engine Piston RingCylinder Liner Conjunction, vol. IMECE2013-62395, 2013. [25] H. Shahmohamadi, M. Mohammadpour, R. Rahmani, H. Rahnejat, C. P. Garner, and S. Howell-Smith, “On the boundary conditions in multi-phase flow through the piston ring-cylinder liner conjunction,” Tribology International, vol. 90, pp. 164–174, 2015. [26] A. Oliva, S. Held, A. Herdt, and G. Wachtmeister, “Numerical simulation of the gas flow through the piston ring pack of an internal combustion engine,” SAE, no. 2015-01-1302, 2015. [27] A. Oliva and S. Held, “Numerical simulation and validation of the flow in the piston ring pack of an internal combustion engine,” Tribology International, 2016. [28] W. Hentschel, A. Grote, and O. Langer, “Measurement of wall
Differences in the strokes could be explained by the piston secondary movement, pressure conditions and direction of movement. It is therefore useful to attain as much measured data as possible. The gas and oil flow in the piston ring pack were calculated with a newly developed CFD-simulation method. The simulation was validated with the measured data from the research engine (ring motion, ring land pressures and blow-by). The simulation results can provide new input for tribological calculations and improve the insights of the gas and oil transport processes in the piston ring pack.
6. Acknowledgements ¨ The research project Kolbenring-Oltransport (IGF-no. 17553) was encouraged by the Federal Ministry for Economic Affairs and Energy on the orders of the German Bundestag with the help of the German Federation of Industrial Research Associations e. V. (AIF) and the Research Association for Combustion Engines e. V. (FVV). The authors thank for the allowance and for the support of the user committee led by the chairman Dr.-Ing. A. Robota (Federal Mogul Burscheid GmbH).
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