Influence of Liner Surface with Parameterized Pit Texture on the Friction Characteristics of Piston Rings

Abstract

The arrangement of a pit-shaped surface texture on the surface of a cylinder liner significantly affects reductions in piston ring friction, and the influence of the structural parameters and spatial distribution on piston ring friction power consumption is unclear. In this paper, the diameter, depth, axial spacing distance, and radial spacing distance of the pits on the inner surface of a cylinder liner were used as variable parameters to process the surface textures of different schemes, and then a friction and wear test was carried out on UMT piston ring–cylinder liner specimens, several texture schemes with the best anti-friction effect were selected, an engine bench test was carried out by comparing these texture schemes with non-texture schemes, and the frictional torque and fuel consumption of the engine were studied at different oil temperatures. The results show that the depth of the pits in the surface texture of a cylinder liner has a greater influence on the friction reduction effect, followed by the radius. The higher the oil temperature in the engine bench test, the greater the impact of the surface texture. The reduction in fuel consumption was greater in the lower-speed region after structuring the textured cylinder liner compared to the non-textured cylinder liner. Specifically, the friction coefficient was mainly affected by the depth of the pits, and the depths of the pits in the texture schemes with good friction reduction effect were all 17–19 μm. The best friction reduction could be achieved when the pit radius is around 50 μm, with little difference in pit depth. When the oil temperature was 95 °C, the average drag torque reduction was about 1.69 Nm; when the oil temperature was 105 °C, the decrease was about 2.54 Nm; and when the oil temperature was 105 °C, the decrease was about 4.53 Nm. After adding the surface texture of the cylinder liner, the fuel consumption rate of the engine equipped with the structured cylinder liner was generally reduced compared with that of the original cylinder liner engine. Among them, the average and subsequent consumption rate of surface assembly scheme 11 decreased the most, with a value of 1.3 g/kwh.

1. Introduction

Diesel engines contain a variety of friction pairs. The operation of each friction pair not only affects the fuel economy and power of the diesel engine but also affects the service life of the engine. Studies have shown that, as it is the most important friction pair in the engine [1,2,3], the friction loss generated by the piston–cylinder liner accounts for 30–35% of the engine friction loss [4,5]; therefore, reducing the friction loss of the piston–cylinder liner friction pair plays an important role in energy saving and emission reduction, as well as in improving engine efficiency. The main functions of the piston ring set are to seal and distribute the lubricating oil on the surface of the cylinder liner, and the movement of the piston ring set within the cylinder liner has a significant effect on minimum oil film thickness, lubricant consumption, and friction losses [6,7].

The working environment of cylinder liners and piston rings is very harsh, the factors to be considered are very complicated, and the lubrication condition changes frequently, so many nonlinear multi-model coupled analyses have gradually become the main methods of studying the friction lubrication characteristics of piston rings and cylinder liners [8,9]. Texturization of surfaces was first proposed by Hamilton et al. in the 1960s [10]. They presented, for the first time, the effect of irregular microcavities between parallel surfaces on hydrodynamic lubrication and verified the potential of the concave/cavity load support mechanism as a source of lubrication using theoretical analysis and experimental verification. For many years, efforts have been made to improve the wear resistance and lubricity of mechanical friction pairs through the application of surface textures. Etsion et al. showed a certain reference value for the study of CLPR surface textures [11], developed a hydrodynamic model of the surface texture CLPR friction system, and also carried out experimental research on some surface texture piston rings to evaluate the impact of surface texture on friction properties. Xu et al. believed that laser processing formed a variety of textures on the surface of boron cast iron cylinder liners, including single grooves, single pits, and composite groove pits [12]. Through the experiment, they concluded that all three textures produced good lubrication. The paper also studied the comparison of isothermal fluids with contact surface lubrication characteristics at different constant temperatures, and the results showed that it is necessary to consider changes in the oil film temperature when studying piston ring–cylinder liner assemblies.

Scholars have carried out many valuable studies on the surface texture of piston rings in the past. Guo et al. [13] showed that the velocity and the applied load of the piston ring are two important factors affecting the temperature of the oil film. A serious temperature rise occurs at the exit of the oil film and the texture area, and the influence of the contact friction heat is not considered in the energy equation. Gu [14], using the average Reynolds equation considering hole correction, the 2D energy equation, and contact friction heat, established a structure surface temperature equation to investigate coating and woven piston ring–cylinder liner mixed lubrication. The results show that low thermal inertial coating, compared to high thermal inertial coating, has lower friction power consumption, and the textured internal fluid may show a local high temperature. Liu et al. [15] showed that the temperature increase caused by frictional heat is significant and greatly affects the tribological characteristics of piston rings. A reasonably designed surface texture can reduce both friction and wear. The influence of pit structure parameters and spatial distribution on the friction power consumption of the piston rings is unclear, and how liner pit texture is affected by oil temperature in engine bench tests is uncertain. This paper firstly processed the surface texture of the different schemes with the diameter, depth, axial spacing, and radial spacing of the inner surface of the cylinder liner as variable parameters and then conducted a friction wear test on the UMT piston ring–cylinder liner, selecting several texture schemes with the best anti-friction effect. Finally, these texture and non-texture schemes were selected for the engine bench test, and the engine friction torque and fuel consumption were studied at various oil temperatures, providing a reference for the parametric design of liner surface texture.

2. Test for Textured Cylinder Liner and Piston Ring Assembly

2.1. Cylinder Liner Surface Texture Parameters

This paper mainly uses the cylinder liners and the first compression ring as examples to model the texture of cylinder liners and the first compression ring, with the specific surface micro-morphology parameters and piston ring–cylinder liner shown in

Table 1. Texture surface structure parameters [16].

Parameter Name	Value
Pit depth hp/μm	5–10
Pit radius rp/μm	25–40
Texture density sp	20–40%
Cylinder liner surface roughness σl/μm	0.4
Piston ring surface roughness σr/μm	0.4
Piston ring barrel height δ/μm	20
Piston ring axial width b/mm	1.5

. The pit-shaped textures were produced using laser processing.

To facilitate the study, certain inhomogeneous factors, such as the lubrication state, load, and variation in wear along the circumferential direction during the reciprocating motion of the piston ring, have been simplified. Since the thickness of the oil film on the contact surface of the piston ring–cylinder liner assembly is significantly smaller compared to the diameter of the cylinder liner, the effect of the curvature of the oil film is ignored during the theoretical analysis, and the cylinder liner is unfolded into a flat surface along the axial direction as shown in Figure 1.

As shown in Figure 1c, the control unit of the pit is a square of length L, and the density S_p of a pit is defined as [17]:

$$S_p=\frac{\mathsf{\pi }{r}_p^2}{L^2}$$

(1)

2.2. Test of the Woven Liner–Piston Ring Assembly UMT

A surface texturing scheme with varying structural parameters is applied to and processed on the inner surface of the cylinder liner. To test the cylinder liner–piston ring assembly with the surface texturing, the UMT-2 multifunctional friction and wear tester is utilized (refer to Figure 2 for equipment illustration), and measurements are taken to determine the average coefficient of friction and friction force. Castrol VECTON high-efficiency diesel lubricant (15W-40) is implemented for the conducted test trials. The key parameters of the UMT trials are listed in

Table 2. Key parameters for UMT tests.

Test Name	Load/kg	Speed/rpm	Trip/mm	Time/min	Test Temperature/°C
Constant parameter friction test	3	1500	3.5	120	40
Variable speed test	3	100–2500	3.5	0.5	40
Variabletemperature test	3	1500	3.5	120	40, 100
Variable load test	1–3	1500	3.5	120	40

.

Friction and wear assessments are conducted on the test specimens of cylinder liners with 18 diverse weave patterns using the UMT-2 multifunctional friction and wear tester (shown in Figure 2).

Table 3. The pit parameters of cylinder liner texture.

Scenario Number	Diameter/μm	Depth/μm	Axial Spacing/μm	Radial Spacing/μm
1	30	5.5	280	585
2	40	7.8	490	493
3	52	12	600	493
4	54	9.86	494	493
5	47	8.05	730	488
6	59	19.5	720	570
7	120	5.5	715	498
8	100	15.2	220	594
9	106	21.9	500	516
10	29	17.2	260	599
11	45	17.0	520	571
12	49	18.9	540	530
13	54	22.8	495	530
14	56	20.2	765	594
15	52	18.1	510	507
16	115	17.1	685	493
17	97	16.1	252	599
18	101	16.9	475	526

shows the 18 weave patterns.

Table 4. ANOVA for the 18 schemes.

	Diameter/μm	Depth/μm	Axial Spacing/μm	Radial Spacing/μm
	30	5.5	280	585
	40	7.8	490	493
	52	12	600	493
	54	9.86	494	493
	47	8.05	730	488
	59	19.5	720	570
	120	5.5	715	498
	100	15.2	220	594
	106	21.9	500	516
	29	17.2	260	599
	45	17	520	571
	49	18.9	540	530
	54	22.8	495	530
	56	20.2	765	594
	52	18.1	510	507
	115	17.1	685	493
	97	16.1	252	599
	101	16.9	475	526
Sum =	1206	269.61	9251	9679
Average =	67	14.978	513.944	537.722
∑iX2ij2 =	96,304	4553.0321	5261,765	5236,565
St. Dev. =	30.197	5.503	172.74	43.353
SS =	15,502	514.724	507,264.944	31,951.611
n =	18	18	18	18
F =	173.94

shows the ANOVA for the 18 schemes, showing F = 173.94 > 2.74, rejecting the hypothesis of correlation that can be tested as independent schemes. The mean friction coefficients for each weave pattern and the corresponding percentage reduction in friction coefficient in comparison to the original machine are demonstrated in Figure 3. The data show that the average reduction in friction coefficient for surface weave schemes 3–6 and 8–15 is positive. This indicates that all weave schemes have some effect on reducing friction compared to the original machine, except for schemes 1, 2, and 7. The depths of schemes 1, 2, and 7 are relatively smaller, resulting in a negative COF reduction rate. The most substantial decrease in friction is observed in weaving patterns 11, 12, and 15, which results in an increase in average friction coefficients of 12%, 9.5%, and 10%, correspondingly, when compared to the friction coefficients of the original machine. It can be inferred from the high-performance lubrication solutions that the pit depth of the weave structure exerts a greater influence on the lubrication properties. Specifically, below 10 µm pit depth, the surface weave structure exhibits an adverse impact on the lubrication properties. The surface weave structure increases the friction coefficient of the friction pair, and the most effective reduction of friction occurs when the depth of the pits is between 17 and 19 µm. The optimum reduction is achieved when the pit depth does not differ much and when the pit radius is 50 µm or less. The best reduction of the friction coefficient is obtained when the pit depth is 17–19 µm.

An investigation of 18 weaving schemes reveals varying levels of friction reduction for piston ring–cylinder liner assemblies. However, some schemes demonstrate a more significant impact on friction reduction. Consequently, the analysis concentrates on the weaving schemes with notably superior friction reduction effects and assesses the impact of these schemes on friction reduction under different boundary conditions.

In order to comprehensively analyze the effects of various weaving techniques on friction reduction in different working conditions and their impact on the average friction coefficient of the piston ring–cylinder liner assembly, we conduct measurements of the average friction coefficient of the piston ring–cylinder liner assembly while varying conditions like engine speed, lubricating oil temperatures, and loads. The specifics of the test method can be found in Figure 4. The mean friction coefficient of the piston ring–cylinder liner assembly is collected and analyzed to produce a diagram highlighting its fluctuation trend under varying working conditions and weaving patterns. Technical abbreviations will be explained upon first use.

In Figure 4, the comparison results of the average friction coefficients between surface texture schemes 10, 11, 12, and 15 and those of the original machine at different engine speeds are provided. Based on these results, it can be inferred that the average friction coefficients of surface texture schemes 10, 11, 12, and 15 exhibit varying decreasing trends at different engine speeds when compared to the average friction coefficients of the original machine. Notably, texture schemes 11 and 15 display the most rapid decrease in average friction coefficients at low engine speeds. The change curves of the four weaving schemes above are all below the average friction curve of the original machine. This indicates that surface weaving technology contributes to reducing friction at different rotational speeds. Texture scheme 15 exhibits a more effective friction reduction effect than the other schemes, whereas scheme 10 has a coefficient of friction smaller than that of the original machine. However, the difference in the values is not significant.

The lubricant is tested at 40 °C and 100 °C using both a non-texture cylinder liner and texture schemes 11, 12, and 15. The average friction coefficients are measured at different temperatures and are presented in Figure 5. The performance of the piston ring–cylinder liner assembly experiences changes with variations in lubricant temperature, signifying that the temperature of the lubricant is a critical factor influencing the reduction of friction in the texture system. The temperature variance within the oil film during its practical operation chiefly arises from the disparity between the produced heat due to the friction motion and the heat dissipated to the cylinder liner and the cooling system. Changes in lubricant temperature not only determine changes in lubricant viscosity but also affect the physical and chemical properties of the friction surface in contact with the material itself. This can be categorized into three specific aspects: firstly, an increase in temperature can cause a change in the phase of the material or particles in the second phase; secondly, the temperature rise can catalyze or accelerate chemical reactions occurring on the material surface; thirdly, the elevated temperature of the oil film generally lowers its viscosity, reducing the friction coefficient.

At an oil film temperature of 40 °C, the woven scheme demonstrates a slightly lower average friction coefficient compared to the non-woven cylinder liner. Scheme 11 exhibits the largest reduction, and it is only in this model that the woven pattern proves to be more efficient in reducing friction at 40 °C. At an oil temperature of 100 °C, the friction reduction effect of texture scheme 15 improves slightly compared to the non-woven and texture schemes 11 and 12. At elevated oil temperatures, scheme 15 proves more appropriate for low-viscosity lubricants.

Figure 6 displays histograms of various texture schemes aimed at reducing the friction coefficients of the piston ring–cylinder liner test specimens subject to different loads. It can be concluded that different texture schemes have varied impacts on friction reduction subject to different loads. At a load of 10 N, the friction reduction effect of the three texture patterns is superior to the other loads. This observation suggests that the texture patterns examined thus far are more efficient in reducing the average friction coefficient at low loads. When the load increases to 40 N, the average friction coefficient for all texture patterns decreases to varying degrees, albeit not significantly. This is primarily due to the contact load determining the stress field distribution in the contact influence area of the friction sub-materials and the size of the contact area. As the load increases, the average friction coefficient increases. With an increased load, the contact area of the friction sub-section experiences elastic–plastic deformation, resulting in an increased friction subcontact area, which leads to a larger friction shear force in turn. It should be noted that, under high loads, materials can experience severe extrusion deformation. This can create stress concentration and exacerbate the development of existing cracks, ultimately leading to fracture or spalling. These issues can harm the reliability of lubrication and shorten the life of the machine.

3. Engine Stand Test

Through the test equipment of the engine reverse towing test, the fuel consumption rate test on the engine piston ring–cylinder liner assembly friction loss change law has a more clear and practical understanding [18,19]. First, the friction torque of unwoven and woven cylinder liners at different oil temperatures is compared through a reverse towing test. And then, based on the results of the reverse towing test, the fuel consumption rate of the woven cylinder liner design is tested at the actual engine operating speed. Analysis and validation of the friction reduction effect of the weaving scheme proposed in the previous experiments are carried out by comparing the trend of changes in fuel consumption rates under different operating conditions.

As shown in Figure 7, they are composed of a diesel engine, CAC AC Dynamometer, fuel consumption meter, and other control systems. The engine inlet temperature is controlled by the inlet air conditioning, the inlet pressure is controlled by the inlet pressure valve, the engine water temperature is controlled by the water constant temperature system, and the fuel temperature is controlled by the fuel consumption instrument. The specific parameters of the diesel engine used for the test are cylinder diameter D126 mm; crank radius R65 mm; rod length L200 mm; and speed n2000 r/min. The lubricant oil number for the test is 15W-40.

Table 5. Instrument and measurement accuracy.

Instrument	Instrument Models	Measuring Range	Measurement Error
Electric dynamometer	JD120	10–1750 r/min,	±1 r/min,
		100–682 Nm,	±0.05%,
Aerometer	ToCeiL20N100	10–−1200 kg/h	±1%
Fuel consumption gauge	PWK100	5–80 kg/h	±0.1%
Engine measurement and control system	FCD-1300	10–1000/5–200 °C 10–1000 kpa/5–200 kpa −100–200 kpa/−100–100 kpa /−10–50 kpa 5–100%/−40–80 °C	±1.5 °C/±1 °C, ±0.1 kpa/0.1 kpa ± 0.1 kpa/0.1 kpa ± 0.05 kpa ± 2%RH/±0.2 °C

shows the instrument model and measurement accuracy used during the test. The reverse drag test is carried out separately by fitting test samples of different woven liners and then measuring the reverse drag torque corresponding to the different rotation speeds.

During the engine reverse towing, mechanical losses include friction losses caused by systems such as piston–liner assemblies, connecting rod assemblies, the crankshaft and camshafts [20]; the pump air loss from the intake and exhaust system; and the mechanical loss caused by the operation of auxiliary components, etc. Because this test only studies friction losses from the piston–liner system, the effect of other friction pairs should be excluded as far as possible; however, due to the limitations of measurement, the friction loss of the connecting rod assembly cannot be measured by disassembly and must be included. Hence, the friction loss measures of the piston block–connecting rod–bearing in this article, and reduction of the frictional loss of the piston–ring liner assembly, are calculated, as shown in

Table 6. Piston ring–cylinder liner assembly friction loss measurement procedure and calculation relationship.

The Part Name	Test Sequence		Comments
	1	2
Original cylinder liner	●		(1) Keep the corresponding part; (2) Removal of air distribution mechanism during testing (including valves, valve springs, rocker arm assembly, push rods, etc.); (3) Remove the auxiliary machinery.
Original piston set	●	●
Original bearing	●	●
Weave the cylinder liners		●
Name of the mechanical loss			Calculation relationship
The difference in mechanical loss before and after the cylinder liner weaving			=1–2

. In summary, the effect of weaving liners on their friction pairs can be measured by examining the change in torque of the engine during the reverse drag test, and the fuel consumption rate of the engine can be seen as the impact of different cylinder liners while ensuring that other variables remain unchanged.

With the weaving solution chosen by the previous friction wear tester, the texture schemes 11, 12, and 15, which perform well, are selected, and a total of four sets of unwoven cylinder liners are tested for reverse towing to study the impact of the different weaving schemes on the friction performance of the engine as a whole. The fuel consumption rate is then tested. We analyze and compare changes in fuel consumption between different texture schemes in the actual operating environment of the engine. The specific bench test process is as follows:

(1)

Break-in phase: Assemble the processed cylinder liner samples to the engine stand, run the engine at idle for 2 h, and then rotate the engine from 800 r/min to gradually increase to 2000 r/min, which increases the load from 20% to 100% in 20% increments each time the speed is adjusted. In this way, the break-in test is completed after a cumulative 10 h run-in, then the full load is carried out at full speed (2000 r/min, 100% load) for 20 h so that the cylinder liner break-in phase is officially completed;

(2)

Reverse drag test: After the break-in test, two backward drag tests are carried out for each set of cylinder liners, increasing the speed from 800 to 2000 and then decreasing it from 2000 to 800. The reverse drag torque and the average value are measured twice separately, with each bank of cylinder liner samples carrying out reverse drag tests at three different oil temperatures (95 °C, 105 °C, and 115 °C);

(3)

Fuel consumption rate test: Assemble unwoven and woven liners to the engine stand for trial, and increase the engine speed in increments of 100 r/min. Gradually increase the engine speed from 600 r/min to 2000 r/min to measure the fuel consumption rate separately at each speed.

In Figure 8, the reverse drag torque for the stock cylinder liner and surface weave cylinder liner conditions at oil temperatures of 95 °C, 105 °C, and 115 °C, are shown. For both original liner and woven-surface liner configurations, the reverse drag torque tends to increase with speed. Unlike for engines with original liner configurations, the reverse towing torque of engines equipped with woven-surface liners shows a decrease as the speed increases. Specifically, the average back-dragging torque reduction of the engine configured with the surface weave 11 cylinder liner is the largest, by about 1.69 N·m, when the oil temperature is 95 °C over the entire measured speed range. At the oil temperature of 105 °C, the average back-dragging torque of the engine fitted with the surface weave 12 cylinder liner decreases the most, by about 2.54 N·m. In comparison, at the oil temperature of 115 °C, the average back-dragging torque of the engine equipped with the surface weave 11 cylinder liner decreases the most, by approximately 4.53 N·m. The specific downward trend is shown in Figure 9. The oil temperature has a greater influence on the lubrication characteristics of the piston ring–liner assembly, and the higher the oil temperature, the greater the impact of the texture of the surface. Analyzing the variation of different oil temperatures in practice is essential for studying the lubrication characteristics of the piston ring–cylinder liner assembly.

Figure 10 shows the change in fuel consumption in the original liner configuration and the woven-surface liner configuration. As can be seen from the figure, overall, engine oil consumption with textured-surface liners is reduced to a certain extent. The decrease in fuel consumption is greater in the lower-speed region, while the surface texture is less effective in reducing engine fuel consumption in the higher-speed region.

The values change as shown in

Table 7. Fuel consumption test results.

Parameters	Fuel Consumption Rate/g/(kw·h)
Original cylinder liner	197.9
Woven cylinder liner 11	196.6
Woven cylinder liner 12	196.9
Woven cylinder liner 15	197.4

. The fuel consumption rate of engines with woven liners is reduced overall compared to those of engines with original cylinder liners, with the largest reduction in the average fuel consumption rate corresponding to surface texture scheme 11, which is 1.3 g/kw·h. The average decrease in fuel consumption rate for the remaining two woven cylinder liners is less than that of scheme 11. This illustrates the positive impact of the textured liner configuration on reducing engine fuel consumption, with scheme 11 having the most significant effect in reducing fuel consumption.

4. Conclusions

Under the dual influence of energy scarcity and environmental pollution, researching the effect of low-friction characteristics of engine piston ring–cylinder liner assemblies is in line with the needs of the automotive industry at the present stage. This paper focuses on the piston ring–cylinder liner system of a diesel engine as a specific research object, and the best friction-reducing scheme among 18 weave design schemes is selected for engine bench testing through the UMT friction and wear test. The results show:

(1)

In the UMT test, the surface texture of the cylinder liner has a positive effect on reducing the average friction coefficient, of which texture schemes 11, 12, and 15 exhibit the best friction reduction effect; thus, it can be derived that the friction coefficient is mainly affected by the depth of the pits, and the depths of the pits in the texture schemes with good friction reduction effect are all 17–19 μm. The best friction reduction can be achieved when the pit radius is around 50 μm, with little difference in pit depth;

(2)

Oil temperature has a strong influence on the lubrication characteristics of the piston ring–cylinder liner assembly, with the effect of surface texture increasing at higher oil temperatures. Specifically, the average back-dragging torque reduction of the engine configured with the surface weave 11 cylinder liner is the largest, by about 1.69 N·m, when the oil temperature is 95 °C over the entire measured speed range. At the oil temperature of 105 °C, the average back-dragging torque of the engine fitted with the surface weave 12 cylinder liner decreases the most, by about 2.54 N·m. In comparison, at the oil temperature of 115 °C, the average back-dragging torque of the engine equipped with the surface weave 11 cylinder liner decreases the most, by approximately 4.53 N·m. Analyzing the variation of different oil temperatures in practice is essential for studying the lubrication characteristics of the piston ring–cylinder liner assembly;

(3)

After adding the surface texture of the cylinder liner, the fuel consumption rate of the engine equipped with the structured cylinder liner is generally reduced compared with that of the original cylinder liner engine. Among them, the average and subsequent consumption rate of surface assembly scheme 11 decreases the most, and the value is 1.3 g/kwh. The introduction of a liner surface texture leads to a more pronounced decrease in fuel consumption in regions with lower rpm. Conversely, its impact on fuel consumption in areas with higher rpm is diminished;

(4)

In the future, the AI algorithm can be used to optimize the casing texture structure parameters to obtain the minimum friction power consumption.