1 Introduction

Tribology contributes to the development of new materials and manufacturing processes, thereby enabling the industry to be more efficient as it extends the lifetimes of elements of machines [1]. Although various improvements are realized, not all areas of engineering benefit equally, as in the case of components that are used in the fluid power systems. In addition, problems that are related to the disposal of materials must be supported by environmental laws that specify how to properly discard worn components of machines. The combination of tribology and environmental care is a vision for the industry of the future [2, 3]. Thus, the use of materials and processes with higher performance and lower impact on the environment should be financially encouraged [3]. Anand et al. [4] describe this combination as “green tribology,” which deals with interacting surfaces and considers energy/environmental sustainability. Green tribology primarily deals with friction and wear, which are principal factors from an energy conservation perspective. Green tribology also involves environmental aspects of lubrication, new materials and alloys, and surface modification techniques [3].

The use of eco-friendly/biodegradable oil as a substitute for mineral oil and the use of alternative surface coating techniques reduces the impact on the environment, which is required by the industry. The combination of biodegradable oils with surfaces that are coated with materials of lower environmental impact is important for the sustainability and growth of the modern industry. Sustainable and efficient hydraulic machine designs consider replacing mineral oils with vegetable oils and synthetic biodegradables. This replacement is due to the high probability of seal rupture and, consequently, oil leakage, which would result in major environmental pollution to water or soils. It is estimated that 50% of all lubricants in the world end up being spilled into the environment through improper disposal, leaks, and accidents. Researchers have discussed the need for new projects for biodegradable oil applications in machines that are using in agriculture, the oil industry (offshore), and hydroelectric plants, among other activities [5,6,7,8].

In recent decades, research on hydraulic oils has been strongly influenced by the need for the development of hydraulic components and systems, and the requirements and their new applications [8]. However, mineral hydraulic oils are subjected to even tighter controls every day due to their environmental and toxicological impacts [9]. Therefore, there is an incentive to use biodegradable fluids, and in some projects, biodegradability has become one of the most important parameters for both fluid choice and lubricant formulation [10]. Three types of base fluids find applications in the formulation of biodegradable, environment-friendly lubricants: mineral oils, vegetable oils, and synthetic lubricants. For applications in hydraulic systems, rapidly biodegradable fluids are classified according to ISO 6743/4 and ISO 15380 as natural esters (HETG type), synthetic esters (HEES type), polyglycols (HEPG type), and polyalphaolefin or hydrocarbon (HEPR type) [11, 12].

Several studies on the use of biodegradable oils that involve bio-based lubricants are presented in the literature, in which results regarding the physicochemical properties and tribological performances were obtained. For the investigating tribological characteristics, various test methods and equipment have been used. However, the main objective remained the same, namely to study the performance of a bio-based lubricant in terms of friction and wear behavior [13]. Majdam et al. [12] evaluated the wear and tear of a hydraulic pump for HEES and HEPR biodegradable hydraulic oils compared to mineral oil. The results demonstrated a lower loss of efficiency in mineral oil testing. However, in terms of eco-friendliness, HEPR had only a 1.03% efficiency loss, while for HEES, it was 7.3%. Another study that was carried by Kučera el al. [14] compared a sliding pair with a B60 bearing and a journal with a contact surface that was made of 16MnCr5 steel. The lubricated tests were conducted using biodegradable oil HEES and mineral oil. In this study, the results demonstrated a satisfactory tribological performance for eco-friendly HEES. However, the mineral oil presented lower temperature at the end of the test. Tkáč et al. [15] examines the use of a biodegradable fluid in the hydraulic system of an agricultural tractor and demonstrates that this fluid had no negative impact on the sealing components of the system.

The function of lubricants is to prevent contact or to reduce friction between sliding surfaces [16]. However, the combination of the properties of biodegradable lubricants that are applied to coated surfaces must be investigated. These contact conditions require research on new materials and surfaces, to assess the synergy or incompatibility that may occur between eco-friendly lubricants and coated surfaces. Therefore, to evaluate the performances of lubricants on these new surfaces, it is necessary to perform sliding tests under various lubrication regimes. Typically, the Stribeck curve and the Hersey number are used for this evaluation [17].

The Stribeck curve is an experimental model that is used to determine the lubrication regimes of a metallic pair under lubricated sliding for evaluating the results of friction and wear [17, 18]. The lubrication regimes may be classified into four types: boundary, mixed, elastohydrodynamic, and hydrodynamic. This approach is linked with the viscosity properties of the lubricant (η0), the speed of rotation of the shaft (Ue), the contact load (P), the root-mean-square roughness of the surfaces (σRMS), and the coefficient of friction [19,20,21]. Typically, the ratio (\(\eta_{0} \cdot U_{e} /P \cdot \sigma_{\text{RMS}}\)) is used in the abscissa axis of the diagram, which is proportional to the thickness of the lubricating film (h), which is calculated based on fluidic lubrication concepts.

The examination of the lubrication regime is a crucial step in a system that utilizes lubricated sliding. The regime determines the efficiency of the formation of the minimum fluid thickness and, consequently, the separation and the friction of the surfaces. The minimum thickness of the lubricant film depends on parameters such as the normal applied load, the sliding speed, the contact geometry, and the rheological properties of the fluid. In scenarios in which there is a physical separation of surfaces, the main property of the fluid is the viscosity [19].

Satisfactory engineering design in the field of fluid power yields improvements in the tribological system of the hydraulic components. These improvements increase the performance and decrease the power consumption of the machines, which reduces the maintenance cost and the frequency of failure in the sealing systems of these components. This concern with design and with the reduction in the energy consumption of the machines directly affects the environment, as it reduces the CO2 emissions into the atmosphere [22].

Due to worldwide concern regarding the disposal of worn-out materials in the environment, several techniques for the deposition of metallic coatings can contribute to the improvement in the surface properties of these mechanical components, which, in some applications, must operate with biodegradable oils. Among the main techniques of deposition, the following are prominent: laser cladding, chemical and physical vapor deposition (CVD/PVD), and high-velocity oxygen fuel (HVOF) thermal spray [23,24,25]. For over 50 years, hard chrome plating has been used; however, this coating has poor tribological properties. In addition, the high levels of hexavalent chromium Cr+6 that are present have high carcinogenic power and result in environmental contamination [26]. The thermally sprayed WC–10Co4Cr alloy—HVOF—realizes satisfactory tribological performance and has lower environmental impact [27,28,29]. In this study, the authors present several results that demonstrate excellent mechanical and tribological characteristics for future applications in hydraulic system components.

In this study, the tribological behaviors of the materials that are used in the manufacture of hydraulic pumps and cylinders were analyzed, which were subjected to lubricated sliding tests. The tests were conducted in a pin-on-disk tribometer, and biodegradable industrial synthetic oils (HEES and HEPR) and traditional mineral oil (HLP) were used as lubricants. An AISI 1045 steel disk that was coated with WC–CoCr alloy and manufactured via thermal spraying—HVOF—and a brass sphere (Cu35Zn) were used to simulate the contacts of the sliding materials. For the evaluation of the lubrication regimes and of the tribological aspects, namely, of the coefficients of friction and wear, mathematical models from the scientific literature and the ASTM G99 standard, which was adapted for lubricated conditions, were used. In addition, measurements of the physical properties and chemical compositions of the lubricants were conducted, along with an analysis of the surfaces of the disk tracks via scanning electron microscopy and energy-dispersive spectrometry (SEM/EDS), to evaluate the friction and wear behaviors and the mechanisms of damage to the surfaces after sliding.

2 Experimental procedure

The analysis of the lubricating oils is conducted with identical tests and with rheological and tribological models that relate the variables of lubrication and the coefficients of friction and wear.

2.1 Pin-on-disk tribometer

A pin-on-disk tribometer, which was adapted from standard ASTM G99-17, was used to simulate the contact sliding of materials that are used in hydraulic components [27, 30]. The pin-on-disk method utilizes a horizontal rotating disk and a pin that is loaded with a calibrated weight. In this study, a 6-mm spherical tip geometry was used as a pin, which supports the applied load (Fn) stationary against a rotating disk, as shown in Fig. 1. This tribometer enables the measurement of the force and the coefficient of friction, which are measured by a load cell. The speed of rotation (ω) and the linear distance that is traveled (S) during the test are monitored by an inductive proximity sensor. The sphere and disk slide inside a reservoir that is designated at “cup,” which contains lubricating oil.

Fig. 1
figure 1

Pin-on-disk tribometer: details of the lubricated sliding test

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The temperature and the humidity were maintained during the tests at 30 ± 3 °C and 40 ± 1.0%, respectively. The experimental parameters are presented in Table 1. For all tests, the tangential speed of the disk was adjusted to 0.4 m/s. Prior to each test, all specimens (disk and sphere) were cleaned in an ultrasonic bath, washed with isopropyl alcohol, and dried. Two disposable syringes were used for the insertion and removal of the lubricant in the cup that was used as a reservoir. At each applied load or total completed time, the used lubricant was removed, and a new lubricant was inserted with a volume of 10 ml.

Table 1 Experimental conditions of the tests
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Two sets of tests were conducted. In the first set of tests (Test 1), the behaviors of the lubrication regimes, the friction coefficient evolution and the minimum film thickness were evaluated using a disk radius of 18 mm for each normal load, as presented in Table 1. Additionally, in these tests, the contamination of the lubricants by solid particles was evaluated based on physicochemical properties, which were measured via inductively coupled plasma spectrometry (ICP). In the second set of tests (Test 2), tests with durations of 1, 2, 3, 4, and 5 h were conducted, thereby resulting in linear distances of 1500 to 7500 m. These tests were conducted using a normal 30-N load on a 20-mm-radius disk. The objective of these long-term tests was to evaluate the mechanisms of damage on pin and disk surfaces and the trends of pin-on-disk wear and friction behaviors, in combination with the use of biodegradable hydraulic lubricants.

2.2 HVOF coating and sliding materials

The disks that were used as test specimens were manufactured from AISI 1045 steel and coated with WC–10Co4Cr alloy using HVOF. As the body, a spherical pin of Cu–35Zn (brass) was used to simulate the materials that are used in the manufacture of hydraulic components. The chemical compositions of these materials are listed in Table 2.

Table 2 Chemical composition (% in mass) of the disk and sphere coating surface
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The tungsten carbide coating was fabricated using a spray gun—HVOF. The main parameters, which are adjusted for deposition by thermal spraying, are presented in Table 3.

Table 3 Parameters for thermal spraying—HVOF
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2.3 Lubricants

Industrial biodegradable hydraulic lubricants with ISO viscosity grade 46 were selected: hydraulic environmental ester oil synthetic (HEES) that was produced from synthetic hydrocarbons, and hydraulic oil environmental polyalphaolefin (HEPR) and related products. In addition, mineral oil hydraulic fluid (HLP) was used in the tests to obtain data for comparison. Table 4 presents the specifications and physical–rheological properties, in which the viscosity for each lubricant was evaluated using a rheometer.

Table 4 Characteristics of the lubricants
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The viscosity behavior as a function of the temperature was determined using the Reynolds mathematical model [31], which is expressed in Eq. 1, where η0 is the dynamic viscosity at atmospheric pressure, TA is the absolute temperature, and “b” and “a” are constants that are obtained empirically from measurements by the rheometer.

$$\eta_{0} = be^{{ - aT_{A} }} \left[ {{\text{Ns}}/{\text{m}}^{2} } \right]$$
(1)

The coefficient of pressure–viscosity (α) is also specified in Table 4. This coefficient affects the formation of the lubrication film and the energy dissipation [32]. The authors show that the pressure–viscosity coefficient of a fluid is directly proportional to the average friction in a sliding contact in the elastohydrodynamic region. This ratio of pressure to viscosity is conventionally expressed by the Barus equation, which is presented as Eq. 2, where ηP is the dynamic viscosity at pressure P, η0 is the dynamic viscosity at atmospheric pressure, and α is the pressure–viscosity coefficient.

$$\eta_{P} = \eta_{0} e^{\alpha P} \left[ {{\text{Ns}}/{\text{m}}^{2} } \right]$$
(2)

The HEES and HEPR ecological lubricant oils were selected according to the characteristics of the application in fluid power systems. HEPR-type lubricants are also classified as ecological fluids and have superior rheological properties over a wide range of temperatures compared to mineral-based oils of the same viscosity grade [33, 34]. Although these ecological products cost approximately seven times more than mineral oils, they are typically used in hydraulic systems that operate at high pressure and at a high temperature. The main advantages of biodegradable fluids are their stability under temperature variation, their satisfactory lubrication properties, and their compatibility with the main types of seals [35].

2.4 Contact conditions and test procedure

The objective of the tests was to simulate the conditions of lubricated sliding contacts of surfaces that are used in components of hydraulic systems. The roughness of the disk and sphere surfaces was within the specifications that were used by the manufacturers of hydraulic cylinders and pumps [36].

In this case, normal loads were applied according in Table 1, and the loads produced mean contact pressures between the ball and the disk of approximately 173 to 1011 MPa. In addition, these seven normal load values, which are denoted as \(F_{n}\) and listed in Table 1, were used to calculate the Hersey number, which is related to the coefficient of friction. Hence, it is possible to identify the lubrication regimes. The mean contact pressure Pmean between the sphere and the disk is calculated via Eq. 3 [21]:

$$P_{\text{mean}} = \frac{{F_{n} }}{{\pi a^{2} }} \left[ {{\text{N}}/{\text{m}}^{2} } \right]$$
(3)

where a is the radius of the contact area that is produced by the contact pressure and Fn is the normal force that is applied in the test. The radius a is determined from the Hertz pressure via Eq. 4.

$$a = \left( {\frac{{3F_{n} R^{\prime } }}{{E^{\prime } }}} \right)^{{\frac{1}{3}}} \left[ {\text{m}} \right]$$
(4)

where \(R^{\prime }\) and \(E^{\prime }\) are the radius of curvature and the equivalent Young’s modulus, respectively. These two quantities can be defined by Eqs. 5 and 6, where \(R_{1X} = R_{1Y} = 3 \times 10^{ - 3} \;{\text{m}}\) (radius of the sphere) and \(R_{2X} = R_{2Y} = \infty\) (radius of the disk).

$$\frac{1}{{R^{\prime } }} = \frac{1}{{R_{X} }} + \frac{1}{{R_{Y} }} = \frac{1}{{R_{1X} }} + \frac{1}{{R_{2X} }} + \frac{1}{{R_{1Y} }} + \frac{1}{{R_{2Y} }} \left[ {\text{m}} \right]$$
(5)
$$E^{\prime } = \frac{{E_{1} E_{2} }}{{E_{2} \left( {1 - v_{1}^{2} } \right) + E_{1} \left( {1 - v_{2}^{2} } \right)}} \left[ {{\text{N}}/{\text{m}}^{2} } \right]$$
(6)

where E1,2 is the Young’s modulus and v1,2 is the Poisson’s coefficient for the sphere and the disk, respectively, and RX and RY are the radii of curvature in the X- and Y-directions of the contact regions, as illustrated in Fig. 2c. The characteristics of the materials that are used as test specimens are presented in Fig. 2a, b and in Table 5.

Fig. 2
figure 2

Macrographs of specimens and the point of contact: a disk and pin (sphere), b a magnified view of the disk surface porosity, and c a schematic diagram of the contact region [37]

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Table 5 Characteristics of specimens—disk and pin
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To determine the thickness of the lubricant film for each loading condition and the lubricants, the elastohydrodynamic regime equation was used [38]. The minimum film thickness (\(h_{{{\text{min}}}}\)) for a circular contact is calculated via Eq. 7:

$$h_{{{\text{min}}}} = 3.63\left( {\frac{{U_{e} \eta_{0} }}{{E^{\prime } R^{\prime } }}} \right)^{0.68} \left( {\alpha E^{\prime } } \right)^{0,49} \left( {\frac{{F_{n} }}{{E^{\prime } R^{\prime 2} }}} \right)^{ - 0.073} \left( {1 - e^{ - 0.68k} } \right) R^{\prime } \left[ {\upmu{\text{m}}} \right]$$
(7)

where Ue is the sliding velocity, η0 is the dynamic viscosity of the lubricant at 30 °C, and k is the ellipticity parameter, which is equal to 1 for a circular contact. By varying the load Fn of sliding, various thicknesses of the lubricating film \(h_{{{\text{min}}}}\) are obtained, which are related to the friction. The behavior of the friction coefficient is represented by the Stribeck curve (Fig. 3), which is related to the lubrication regime via the Hersey number and is calculated via Eq. 8 [17]. The RMS roughness values, which depend on the equivalent amplitude of roughness (σRMS), are calculated from \(S_{q1}^{2}\) and \(S_{q2}^{2}\), which are the average roughness of the disk and sphere, respectively, via Eq. 9.

Fig. 3
figure 3

Stribeck curve: the friction coefficient μ versus the Hersey number Hs

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A total of three disks and 36 spheres were used for the sliding tests. In Test 1 (radius of 18 mm), one disk and seven spheres were used for each lubricant, and the tests were repeated two times for each load to ensure repeatability. In Test 2, the disk of the previous test (radius of 20 mm) was used, but with a set of five new spheres for each lubricant, and the wear of the sphere was evaluated every hour of test. The roughness (Ra) of the surfaces of the disks was measured using a roughness tester and the roughness of the sphere was defined by ISO 3290-1. In addition, a microhardness tester was used to measure the hardness of the specimens.

$$Hs = \frac{{\eta_{0} U_{e} }}{{P_{\text{mean}} \sigma_{\text{RMS}} }} \;\left[ {\text{non-dimensional}} \right]$$
(8)
$$\sigma_{\text{RMS}} = \sqrt {S_{q1}^{2} + S_{q2}^{2} } \left[ {\upmu{\text{m}}} \right]$$
(9)

The film parameter (λ) relates the film thickness and the RMS value of the roughness amplitude, namely σRMS. λ is also known as the “Lambda factor,” and its value depends on the lubrication regime and the Stribeck curve in the range of values that were obtained experimentally. The parameter of the film is calculated via Eq. 10:

$$\lambda = \frac{{h_{{{\text{min}}}} }}{{\sigma_{\text{RMS}} }}\; \left[ {\text{non-dimensional}} \right]$$
(10)

2.5 Determination of the coefficients of friction and wear

The coefficient of friction was calculated from the measurement of the instantaneous friction force that was provided by the tribometer, which was divided by the normal force that was exerted by the pin, using the concepts that were proposed by Amontons and Coulomb [39]. The diameter of the wear scar (WSD), which is denoted as d, was measured using an optical microscope in the directions of the ordinate and abscissa at the end of each test. To calculate the volume that was removed from the sphere surface, it was necessary to calculate the height that was removed from the volume, namely, h, via Eq. 11, from the radius r of the sphere and the diameter d of the scar that is shown in Fig. 4. The removed volume, which is denoted as Q, is calculated via Eq. 12, which depends on the hemispherical geometry [40]:

$$h = r - \left( {r^{2} - \left( {\frac{d}{2}} \right)^{2} } \right)^{{\frac{1}{2}}} \;\left[ {\text{m}} \right]$$
(11)
$$Q = \pi h^{2} \left( {r - \frac{h}{3}} \right)\; \left[ {{\text{m}}^{3} } \right]$$
(12)

where d is the wear scar diameter, r is the radius of the half-sphere, and h is the height of the wear volume. To determine the coefficient of wear, namely K, the wear model that was proposed by Archard is used, as expressed in Eq. 13 [42].

$$\mathop \sum \limits_{i = 1}^{n} Q_{i} = K\left( {\mathop \sum \limits_{i = 1}^{n} F_{{n_{i} }} S_{i} } \right) \left[ {{\text{mm}}^{3} } \right]$$
(13)

where Si is the sliding distance of each test in m, K is the wear coefficient in mm3/Nm, and \(F_{{n_{i} }}\) is the normal load in N. The removed volume Qi is cumulative over all tests and is represented by a linear equation.

Fig. 4
figure 4

Scar wear and parameters for calculating the volume that was removed from the sphere (Li et al. [41])

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After the lubricated sliding tests were conducted, the track surface failure mechanisms were evaluated via scanning electron microscopy (SEM) with dispersive energy spectrometry (EDS) analyses.

3 Results and analysis

An analysis was conducted for each test condition to evaluate the lubrication regimes of each lubricant and the damages on the surfaces of the disks and of the spheres that were used in the tests. Thus, it is possible to simulate the sliding of materials that are commonly used on the surfaces of hydraulic components. With these results, it was possible to evaluate the tribological behaviors of the biodegradable hydraulic oils.

3.1 Viscosity

To determine the lubrication regimes and the film thicknesses, it was necessary to determine the viscosity behavior of the lubricants, mainly at the mean temperature of 30 ± 3 °C of the test. The coefficients, namely a and b, of the Reynolds equation (Eq. 1), its experimental correlation, and the dynamic viscosity that were obtained for the exponential viscosity versus temperature variation are listed in Table 6. These coefficients were obtained in the range of 10 to 70 °C.

Table 6 Fitting parameters according to the Reynolds equation for the viscosity at 30 °C
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According to the viscosity data that were obtained in the laboratory after the tests, the values of coefficients a and b are lower for the biodegradable hydraulic oil (HEPR); hence, this fluid is of lower viscosity. In addition, higher similarity of this property is observed for HEES and HLP oils.

3.2 Lubrication regimes and film thickness

The performance of the lubricant in various lubrication regimes is typically characterized using the Stribeck curve, which relates the coefficient of friction μ with the variables of contact pressure, sliding velocity, dynamic viscosity, and RMS roughness using the Hersey number, which is denoted as Hs [17, 43]. In the tests, Stribeck curves were obtained for various loads and a single sliding speed according to the parameters that are specified in Table 1 (Test 1) via Eq. 8. The tests were conducted at a total sliding distance of 125 m for each load. Figure 5 presents the results of the lubrication regimes for the sliding of the brass/WC–CoCr surfaces.

Fig. 5
figure 5

Stribeck curve: The coefficient of friction versus the Hersey number for each lubricant and applied load

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According to Fig. 5, the coefficient of friction varied from approximately 0.02 to 0.08 in the range of Hs between 1.38E − 8 and 1.17E − 6. The lowest values of the coefficient of friction for the lubricants were between 0.02 and 0.03, and this region is characterized by a change in the regime of elastohydrodynamic lubrication. For a higher coefficient of friction and a low Hersey number (smaller than 1.38E − 8), the regime is mixed lubrication. For higher values of the coefficient of friction, namely, greater than 3.7E − 7 Hs, the trend is to modify to the hydrodynamic regime, but with the challenge that a punctual contact of the sphere would have to form a lubricating film. Under this condition, the roughness without deformation by the medium pressure predominates, which is a requirement that characterizes the hydrodynamic regime.

The behaviors of the lubricating oils with decreasing loads and increasing Hersey number were analyzed and compared, and the following results were obtained: In the range that was defined as “Region S,” the value of the coefficient of friction μ from point 1 to point 2 decreased from 0.073 to 0.022 with biodegradable oil HEES. For the mineral oil, namely, HLP (Region P), the coefficient of friction from point 3 to point 4 had a variation of 0.060 to 0.019. In the region R that was defined for biodegradable oil HEPR, the coefficients of friction ranged from 0.055 to 0.030 between points 5 and 6 of Fig. 5. By evaluating the region of mixed lubrication and elastohydrodynamics, it was observed that during the tests, oil HEES corresponded to higher values of the coefficient of friction for most of the evaluated loads, compared to the other lubricants. In the hydrodynamic lubrication regime for the higher Hersey value of the HEES oil, which corresponds to a load of 0.3 N, the highest value of the coefficient of friction is observed. While the HLP oil corresponds to lower loads and higher Hersey numbers, the coefficient of friction tends to maintain the same value because the film thickness remains the same. This phenomenon of lubrication, which is known as starvation, can occur in the regime of hydrodynamic lubrication [44]. The authors further state that the lubricant film is not only related to the pressure, speed and viscosity as independent parameters but also to the level of oil filling during sliding of the surfaces.

The friction depends on the film thickness, the roughness contact, the temperature increase and the structural deformations of the components [45]. The film thickness is calculated via Eq. 7. The following values were obtained for the dynamic viscosity (η0) and the viscosity-pressure coefficient (α): \(\eta_{0} = 0.060 \;{\text{and}}\; \alpha = 1.38 \times 10^{ - 8}\) for HEES oil, \(\eta_{0} = 0.057\; {\text{and }} \;\alpha = 1.82 \times 10^{ - 8}\) for HLP oil, and \(\eta_{0} = 0.051\; {\text{and }} \;\alpha = 1.58 \times 10^{ - 8}\) for HEPR oil. Therefore, the values of the minimum film thickness of HEES oil and HEPR are lower than that of HLP. The viscosity and viscosity–pressure coefficients values affect the film thickness. These properties depend on the molecular structure of each lubricant. Figure 6 shows comparison of the calculated film heights for the various mean contact pressures (\(P_{\text{mean}}\)) that were applied in the tests, and the coefficient of friction and the film parameter λ as functions of the Hersey number.

Fig. 6
figure 6

Lubrication film evaluation. a The film thickness with respect to the mean contact pressure and b the film parameter and coefficient of friction as functions of the Hersey number

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Figure 6a shows that the result of the film thickness as a function of the mean contact pressure has a logarithmic characteristic that was modeled by the experimental data, which corresponds to a curve quality estimate of over 99%. Figure 6b shows the relation of the Stribeck curve with the film parameter,\(\lambda\), whose units are represented in the secondary axis of the graph. In this result, differences of 9 and 13% of the thickness of the HLP lubricant film for the HEES and HEPR lubricants, respectively, were identified.

The values of the friction coefficient and the film parameter (λ) as functions of the Hersey number are presented in Table 7, which analyzes the behavior of the coefficient of friction and the modification in the lubrication regime, along with the values of the film parameter for the ecological and mineral lubricants.

Table 7 Coefficient of friction and film parameter (λ) as functions of the Hersey number for each lubricant
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The film parameter expresses the severity of the contact roughness during the sliding. In previous lubricated sliding tests of a sphere on a disk that was coated with three metallurgical powders, values of \(\lambda \cong 0.5\) for tests with 0.5 m/s speed and \(\lambda \cong 2.2\) for tests with 4 m/s speed were obtained [41]. It was defined that if 1 > \(\lambda\) > 3, there is a mixed lubrication regime, and the lubrication limit would be λ = 0.5.

The divisions of the mixed, elastohydrodynamic and hydrodynamic lubrication regimes are identified based on the changes in the values of the coefficient of friction, which are mainly near the upwardly concave point, according to the tests that were conducted [45]. However, other research indicates that these results would have acceptable precision for light and moderate loads, but not for high loads [46]. In the sliding test, the lowest coefficient of friction is found in the region of concavity that is identified in Table 7. The friction coefficient for HEES biodegradable oil is \(\mu_{{{\text{HEES min}}}} = 0.022,\) which corresponds to the film parameter value of \(\lambda_{\text{HEESi}} = 0.29\). The friction coefficient of the mineral oil HLP, namely, \(\mu_{{{\text{HLP min}}}} = 0.019,\) corresponds to a film parameter value of \(\lambda_{\text{HLPi}} = 0.41\). For biodegradable oil HEPR, the coefficient of friction was \(\mu_{{{\text{HEPR min}}}} = 0.029\), which corresponds to \(\lambda_{\text{HEPRi}} = 0.35\). The Stribeck curve shows that in the hydrodynamic regime, the HEPR oil had the lowest coefficient of friction, which is recommended for moderate loads and medium speeds. However, HLP and HEES outperform it in the elastohydrodynamic regime, as they are more suitable for higher loads and speeds than HEPR.

The values of the film parameter λ that are presented in Table 7 do not accord with previous research [41]; this is because the WC–10Co4Cr coating material had high porosity, as shown in Fig. 2b, and a substantial roughness difference due to the surface preparation process of the disk. This type of surface has small reservoirs for lubricating oil that are not considered in the film thickness calculations; however, it performs well for lubrication and yields smaller values for λ.

The average roughness of the brass sphere was \(S_{{q_{\text{pin}} }} = 0.015 \;\upmu{\text{m}}\), whereas the roughness of the disk that was used with the HEES lubricant was \(S_{{q_{\text{HEES}} }} = 0.135 \;\upmu{\text{m}}\); hence, the roughness of the disk was nine times larger than of the sphere. The disk that was used for testing with the HLP lubricating oil has a HLP roughness of \(S_{{q_{\text{HLP}} }} = 0.121 \;\upmu{\text{m}}\), and the sphere maintains the same value, in which case the roughness of the disk is approximately 8 times greater. For the HEPR oil, the value of the roughness of the disk, namely \(S_{{q_{\text{HEPR}} }} = 0.105 \;\upmu{\text{m}}\), corresponds to seven times the difference of the sphere. The roughness values influence substantially the calculation of the film parameter \(\lambda = h_{{{\text{min}}}} /\sigma_{\text{RMS}}\). Since the disk roughness is approximately 7 to 9 times greater than the sphere, the high value of roughness in the denominator promotes low values of the film parameter. The results of other studies that were conducted on a lubricated test machine demonstrate that for values of λ > 5, the friction does not increase indefinitely; with higher speeds, the friction begins to decrease again [47]. This is because with a high shear rate, the lubricant begins to heat, which decreases the viscosity and, therefore, causes a decrease in the friction. The friction is not affected by the longitudinal roughness but depends strongly on the peak-to-valley height, namely, the equivalent roughness [48]. In Fig. 5, at the last point for the HLP oil, which corresponds to the highest Hersey number value, stabilizing behavior of the coefficient of friction was also observed, which was due to the lubrication-starved conditions.

Other studies show that high pressure, antifriction, and anti-wear additives substantially decrease the value of the coefficient of friction and, consequently, the value of the coefficient of wear [49]. In this study, the authors identified high concentrations of zinc, phosphorus, and molybdenum, which are usually used in lubricant formulations as extreme-pressure and anti-wear additives.

Table 8 lists the concentrations of elements in lubricants HEES, HLP, and HEPR, which were determined via inductively coupled plasma atomic emission spectrometry (ICP).

Table 8 Concentrations of the elements that were used as additives in the lubricants
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According to the concentrations of the elements that were used as additives that are presented in Table 8, the synthetic biodegradable oils, namely, HEES and HEPR, have lower concentrations of Zn and P, which are used as anti-wear additives. The results demonstrated that there was 29.2 times more Zn in HLP than in HEES oil and 11.8 times more in HEPR. For phosphorus (P), the values were 3.2 times that in mineral oil for HEES and 2.3 times for HEPR. These higher concentrations of zinc and phosphorus in the mineral oil increase its capacity to withstand higher loads without affecting the contact of the surfaces, which, consequently, reduces the coefficient of friction compared to the evaluated biodegradable oils. The lubricants are often mixed with zinc dialkyldithiophosphate (ZDDP) as a multifunctional additive [49]. Zn and P adhere to the surface of the steel and protect this surface against abrasive and adhesive wear mechanisms, thereby reducing the friction and wear during the sliding. Table 8 also presents the differences in sulfur concentrations, which, in combination with other elements, is used as an extreme-pressure additive. The lowest concentration was identified for HEES and the highest for HEPR oil, thereby revealing the effect of this element on the stabilization of the friction coefficient for the HEPR oil, as shown in Fig. 6b. The results of other studies support the effect of sulfur as a wear-inhibiting element and the reduction in the coefficient of friction [50]. In this study, the authors evaluate the wear and friction behaviors in the sliding of spheres that were manufactured from AIS 52100 and steel disks.

3.3 Results of friction and wear mechanisms—Test 1

The wear and friction tests were conducted with loads of 0.3, 1, 4, 10, 30, 50, and 60 N, at a total distance of 125 m. To evaluate the wear of spheres after lubricated sliding, it was necessary to measure the wear scar diameter (d) at the end of each test. From the measurements of the scar diameter, the removed height (h) of the pin (sphere) was determined via Eq. 11, from which the volume loss of the sphere (Q) was calculated via Eq. 12. Figure 7a shows nonlinear behavior for the wear scar (WSD) for each load and a scar morphology that indicates higher wear for the spheres with HEES biodegradable oil. The differences in the wear scars are visible after sliding, as shown in Fig. 7b–d.

Fig. 7
figure 7

Wear of the spheres: a the evolution of a wear scar (WSD) for each load and b–d sliding with HEES, HLP and HEPR, respectively, for the load of 60 N

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The largest difference was observed for the WSD with the load of 60 N. With this load, the average WSD with HEES was 1.301 mm (Fig. 7b), compared to 0.996 mm for HLP (Fig. 7c)) and 0.696 mm for HEPR (Fig. 7d). According to the values of the average friction coefficient for all loads during sliding, the mineral oil presented a mean value of the friction coefficient of \(\mu_{\text{HLP}} = 0.044\), compared to \(\mu_{\text{HEES}} = 0.054\) and \(\mu_{\text{HEPR}} = 0.043\); hence, the biodegradable oil HEES has difficulty maintaining a lubricant film between the sliding materials.

From the data on the removed and accumulated Q volumes of the sphere as functions of the load and the distance traveled, the dimensional wear coefficient K is calculated via Eq. 13. The wear behavior of the spheres that were used as pins in the sliding test for each lubricant is presented in Fig. 8.

Fig. 8
figure 8

Results for the wear coefficient K in mm3/N m for each lubricant

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The coefficient of wear (K), which was obtained via Eq. 13, for the sphere of Cu-35Zn shows linear behavior for the three lubricants. For the HEES biodegradable oil, the value is \(K_{\text{HEES}} = 6.90 \times 10^{ - 5} \;{\text{mm}}^{3} \;{\text{m}}/{\text{N}}\) with a correlation coefficient of R2 = 99.7%. With the mineral oil, \(K_{\text{HLP}} = 2.12 \times 10^{ - 5} \;{\text{mm}}^{3} \;{\text{m}}/{\text{N}}\) with R2 = 99.8%. For the biodegradable HEPR oil, \(K_{\text{HEPR}} = 8.62 \times 10^{ - 6} {\text{mm}}^{3} \;{\text{m}}/{\text{N}}\) with a correlation of 99.6%; hence, in all cases, the quality of the data that were used in the linear regression is satisfactory. The differences in wear among the lubricants are substantial, namely, the removed volume accumulated and Qi with use of the HEES biodegradable oil was 1.34 mm3, compared to 0.4 mm3 with the mineral oil. Comparing the HEPR and the HLP, the removed volume of the biodegradable oil was lower, with a value of 0.16 mm3. In the end, the sphere wear coefficient with the HEES biodegradable oil was 3.2 times higher than that with HLP and 8.3 times that with HEPR.

The lubrication regimes and partial wears in the short-term tests are compared in Fig. 9 for sliding with the HEES, HLP and HEPR oils.

Fig. 9
figure 9

Comparison of the coefficient of friction and the removed volume from the sphere versus the Hersey number. a HEES—esters, b HLP—mineral oil, and c HEPR—hydrocarbon

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The Hersey number is compared with the coefficient of friction and removed volume from each test. For all lubricants, in low loads and in the elastohydrodynamic regime, small and negligible wear occurred prior to satisfying the concavity condition of the Stribeck curve. In this regime, there is no contact between the asperities of the surfaces. If the Stribeck curve satisfies the concavity condition, small but noticeable wear occurs. Hence, wear could be avoided with larger values of the Hersey number than the value that defines the concavity condition. For the HEES oil, the lowest coefficient of friction was attained with the load of 4 N. However, with smaller loads (greater than 1 N), a lower coefficient of friction and likely less severe wear could be realized, as detected in the boundary lubrication regime. Additionally, under the concavity condition, the lowest coefficient of friction with the HEPR oil was with attained the load of 4 N; however, with this oil, lower wear was realized compared to the HLP mineral oil and the biodegradable HEES oil. In all tests, it was observed that the HEPR oil corresponded to less variation of the friction coefficient in the lubrication regime. After leaving the range that is defined by the concavity condition, the wear differs substantially among the oils; it is mild for HLP (mineral) and HEPR (hydrocarbons) and highly severe for the biodegradable HEES (ester) oil. In the mixed lubrication regime, contacts will occur between the highest peaks of the roughness. For all sliding conditions of the lubricants, the additives have affected HEPR and HLP strongly and positively due to the high concentrations of Zn and P, which are specified in Table 8. At the point of the largest load of 60 N, which corresponds to the smallest Hersey number for HLP and HEPR (Fig. 10b, c), the wear was slightly reduced compared with the load of 50 N. This is due to a reaction of the anti-wear and extreme-pressure additives that operate optimally under these conditions.

Fig. 10
figure 10

Results of the coefficient of friction at various test times: a 1 h, b 2 h, c 3 h, d 4 h, and e 5 h, and f the trend of the coefficient of average friction at each test time

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In other lubricated sliding tests of a copper block under an SAE 52100 steel wheel, with elastohydrodynamic lubrication, the coefficient of friction, the wear rate, and the temperature exhibited low and constant values [51]. In contrast, mixed lubrication is characterized by a higher coefficient of friction. Consequently, the wear would occur due to the lack of lubricant film, thereby causing scratching, removal, deposition, and crushing of particles in the wear track; if the material is of high ductility, the wear would be caused by successive passages of the sphere on the track.

After the analysis of the wear on the sphere and disks, the results regarding the contamination of the lubricants, which was due to the wear of the bronze sphere, are presented, which were obtained according to ASTM D5185. Additionally, particle count results were obtained for each lubricant according to ISO 4406. Table 9 presents the results for each hydraulic oil before and after the wear tests.

Table 9 Amounts of copper that were detected in the lubricants in parts per million (ppm) via spectrochemical analysis and particle counting according to ISO 4406 before and after the wear tests
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Table 9 presents the concentration of copper particles for each lubricant after the sliding tests. The results demonstrate that the number of particles that resulted from the wear of the sphere after the tests was higher for the HEES biodegradable oil, which corresponds to an increase of approximately 91 times, compared to 8.6 and 3.1 times for HLP and HEPR, respectively. Previous studies identify the elements (Fe, Cu, and SiO2) that are the main contaminants in a lubrication system and demonstrate that high concentrations of these elements increase the coefficient of friction and the wear of the surfaces in contact [52]. Table 9 also supports the evolution of the number of solid particles, in which the mineral oil presented with lower contamination than the biodegradable lubricating oils, which is due to the changes of the three codes of the standard. Particle quantification with this standard is typically conducted with sizes ≥ 4, 6 and 14 μm, which are identified by codes 9 to 28. In the results that were obtained after the tests, greater contamination in the biodegradable oils is observed since the codes underwent greater modifications, compared to without oil use. The concentration of copper as a contaminant in biodegradable oils also substantially impacts the increase in the viscosity, the oxidation, and the aging of the lubricating oil [53].

3.4 Results of friction and wear mechanisms—Test 2

In most studies on the tribological conditions of lubricated surfaces, short-term tests are conducted. However, via short-term tests, it is not possible to identify a trend in the coefficient of friction or to predict the damage that was caused to the surface by its modification. In this study, the results of the friction coefficient throughout the test for each lubricant are primarily presented separately for each hour of the test. In the tests, three disks were coated in WC–CoCr, namely one for each lubricant. The objective of this test was to simulate the sliding of hydraulic components that are not replaced after a period of use and to evaluate their behavior under critical conditions. Figure 10 presents the results of the coefficient of friction for long tests with a load of 30 N. Under these load and speed conditions, the mixed lubrication regime was observed, as shown in Fig. 5.

The friction force signals that were provided by the load cell have been filtered at low-pass frequencies so that the electrical noise and machine vibrations do not affect the frictional forces. If the normal force is constant and the friction force consists of only force values, the coefficient of friction has small amplitudes under stable and differentiated conditions and under floating conditions.

In all tests, the coefficient of friction for the HLP mineral hydraulic oil is the most stable, followed by HEES oil and HEPR oil. The coefficient of friction, which is plotted in Fig. 10, is analyzed as follows for each test time.

In the tests that were conducted for a duration of 3600 s, the average values of the friction coefficient for the lubricating oils were \(\mu_{\text{HEES}} = 0.050, \mu_{\text{HLP}} = 0.060 \;{\text{and}}\; \mu_{\text{HEPR}} = 0.032\). In this test, low instability for the HEPR was observed from 2000 s. In the second test, which was 7200 s in duration, a higher average value of the coefficient of friction for HEES oil was observed, along with instability of HEPR, compared to mineral oil HLP. The values of the average coefficient of friction were \(\mu_{\text{HEES}} = 0.078, \;\mu_{\text{HLP}} = 0.056\;{\text{and}}\; \mu_{\text{HEPR}} = 0.041\). It is also found that from this test that the stick–slip behavior becomes more evident throughout the HEPR oil sliding test. The frictional force changes with the elasticity of the dynamic system and the speed, which would be close to the critical speed of the mass–spring microsystem. This change occurs due to the transition from mixed lubrication to boundary lubrication, which influences the friction and changes the speed from highest to lowest, thereby approaching dry friction [54]. According to previous research, the maximum amplitude of the stick–slip is proportional to the difference between the static and kinetic friction forces [55]. However, in the experimental results that are presented in Fig. 10 for HEPR oil, higher values were observed for the coefficient of static friction. In the 3-h (10,800 s) tests, the mean coefficient of friction for the HEES was higher than for the other lubricating oils. In this test, the instability for the HEPR oil started at approximately 1800 s, and this behavior was maintained until the end of the test, with an increasing trend throughout the test. In these tests, the average COF values for the oils were \(\mu_{\text{HEES}} = 0.081, \;\mu_{\text{HLP}} = 0.069 \;{\text{and}}\; \mu_{\text{HEPR}} = 0.079\). In the fourth test, which was conducted for a duration of 14,400 s, the HLP oil again presents better stabilization, and an increasing trend is observed for the HEES oil, in addition to a sudden increase in the coefficient of friction at approximately 11,000 s for this fluid. This result is explained by the adhesion mechanism on the surface of the disk, which demonstrates the low performance of the friction and wear additives for this biodegradable ester-based oil. The adhesion mechanisms result in metal-to-metal contact and likely increase the coefficient of friction. As the wear is developing, debris is released and slightly increases the roughness of the surfaces in an indirect way; a transitorily decreasing in the film parameter (λ) is observed [41]. The results of the tests that were conducted at 5 h were similar to the previous results. In these tests, the average COF values that were obtained for the three oils were \(\mu_{\text{HEES}} = 0.083, \;\mu_{\text{HLP}} = 0.063\;{\text{and}}\; \mu_{\text{HEPR}} = 0.047\). In all the tests, the values that were obtained for the coefficient of friction for the HEPR oil were lower than those for the other lubricants. Figure 10f shows the behavioral trends of the average friction coefficients for each test time. The linear regression yields coefficients of 0.0112, 0.0015, and 0.0076 for HEES, HLP, and HEPR, respectively. According to these values, the coefficient of friction for the biodegradable HEES oil is the highest, followed by that for the biodegradable HEPR oil and, finally, that for the HLP mineral oil, thereby demonstrating the performances of the lubricants during the sliding of the WC–CoCr/Cu–Zn tribological pair.

The long-term tests were used to evaluate the volume that was removed from the sphere via Eqs. (11) and (12) and to obtain the coefficient of wear via Eq. (13) for each lubricant. Figure 11 shows the comparison of the accumulated wear coefficients for the 5 evaluated times.

Fig. 11
figure 11

Wear coefficient K in mm3/N m for each lubricant with a one-hour interval between tests

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The comparison of the performances of the lubricants with respect to the wear resistance shows the following: The HPL oil showed a lower coefficient of wear, which was on the order of \(K_{\text{HLP}} = 7.98 \times 10^{ - 7} \;{\text{mm}}^{3} \;{\text{m}}/{\text{N}}\), and its removed volume was lower compared to HEES. The linear correlation of 95.3% for HLP demonstrates that its trend is to maintain its wear resistance. The HEPR oil performed well with a wear coefficient of \(K_{\text{HEPR}} = 1.81 \times 10^{ - 6} \;{\text{mm}}^{3} \;{\text{m}}/{\text{N}}\), and the ratio of the HEPR wear coefficients with HLP is 2.27. The HEES oil did not provide satisfactory wear resistance during the sliding of materials. The wear coefficient that was obtained for the sphere while using this lubricant was \(K_{\text{HEES}} = 1.29 \times 10^{ - 5} \;{\text{mm}}^{3} \;{\text{m}}/{\text{N}}\), and its distance, which is plotted in Fig. 11, shows that the wear was greater and evident. The ratio of the HEES wear coefficients with HLP is 16.7 and that with HEPR is 7.13.

The performance results of the lubricating oils with respect to wear and friction are compared in Fig. 12. These results consist of the removed volume for each test time and the behavior of the coefficient of friction.

Fig. 12
figure 12

Relationship between the removed volume Qin mm3 and the coefficient of friction µ versus the test time t in hours

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According to Fig. 12, the trends are not linear in any case and the behavior can be summarized as follows: For the first three hours, there is a tendency for increased friction and wear for all lubricants. The increase in wear in the first three hours is not proportional, and although the coefficient of friction decreased slightly in the second hour for the HLP oil, its increase in the third hour is approximated by the average of the other oils. For the HEES oil, in the third hour, a slight decrease was observed. In the fourth hour, the coefficient of friction was higher for the HEES and HEPR oils, whereas that for the HLP oil decreased slightly. The greatest wear of HEPR occurs the fourth hour. The coefficient of friction of the HEES oil was much higher than those of the other oils due to the abnormal behavior of the coefficient of friction, which is shown in Fig. 10d, which reaches the value of the coefficient of dry friction. This is due to the decrease in lubrication at this time. In the fifth hour, the coefficients of friction decrease with all lubricants; the decreases are substantial for HEES and HEPR but almost imperceptible for HLP. Although the wear of the spheres with the use of HEES is higher than for the other lubricants, the trend shows a decrease in the wear rate. The trend of increasing wear was prevented by the changes in the coefficient of friction, as shown in Fig. 10e. The changes in friction and wear for the brass in the limit lubrication regime occur due to a strong plastic deformation of the debris, with intragranular sliding in the α phase, and appear with localized deformation in shear bands, with an increase in hardness after the tests [51]. According to the experimental results of Fig. 12, the changes in the coefficient of friction and wear would be affected by the detachment of the brass over the disk track, thereby filling a portion of the pores of the disk coating. This third body, which is added instantly by debris and crushing with adhesion to the disk surface, has modified the initial characteristics of the surfaces, thereby reducing the friction and wear. With the increase in the number of passes or repeated sliding, the coefficient of friction for copper alloys gradually decreases [56]. Several mechanisms of wear have been evaluated and discussed for mechanical systems, such as abrasive wear, adhesive, and fatigue. In the tests that were conducted in this study, mostly in the mixed lubrication regime, abrasive wear and adhesive mechanisms were identified after the sliding on the tracks. Figures 13, 14, and 15 present the micrographics that were obtained via SEM of the worn surfaces of the disk after the Cu–Zn alloy sliding, in addition to the chemical compositions that were made by EDS outside and within the wear track.

Fig. 13
figure 13

SEM-EDS micrograph and chemical compositions of a WC–CoCr surface that was lubricated with HEES

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Figure 13a shows a micrograph of the wear track that was produced by the sliding of the Cu–Zn sphere against the disk that was made of WC–CoCr and lubricated with HEES hydraulic oil. In the magnified view of the track surface, a material that has been deposited on the disk is identified in darker color, thereby demonstrating the adhesive wear mechanism, which was subsequently supported by the chemical composition. Figure 13a still shows grooves that are aligned in the direction of sliding, which are characteristic of an abrasive wear mechanism. Then, the adhesion of the sphere material on the disk was supported by the analyses that were conducted via EDS (Fig. 13b). High concentrations of copper (63.86%) and zinc (29.33%) elements from the sphere material, which were higher than those of the other elements of the WC–CoCr alloy, were identified. In Fig. 13c, the chemical elements that are present on the surface of the disk outside the wear track (P1.2) are identified for comparison with the worn surface (P1.1), for evaluating the mechanisms. The adhesive wear occurs with the sliding of the metals when the surfaces are in contact [57]. During the sliding, it is possible that microsoldering occurs between the metals, with a subsequent detachment of these materials. The more compatible the metals, the more severe the adhesive wear. Metals are considered compatible when their weldability is between 0.1 and 1%. If it is below 0.1%, the metals are considered partially incompatible. If it is outside this range, the metals are characterized as not compatible. In the tests, the predominant material on the disk surface is tungsten (W—70.2%), whereas in the sphere, it is copper (Cu—62.3%). The tungsten carbide alloy and brass that were used in the tests are partially compatible; therefore, adhesive and abrasive wear occur [58].

Figure 14a presents the same information for the lubricated sliding but with the HLP oil. In this result, the mechanism of abrasive wear on the track was more evident after the tests. In the magnified view of the track region, it is possible to better identify the alignment direction of the grooves and the mechanism of surface damage, such as chipping failure in the coating. Figure 14b presents the amounts of Cu (0.63%) and Zn (0.83%) on the surface, which are much lower than in the wear track with HEES oil; hence, HLP mineral oil presents superior tribological performance for surfaces. Additionally, in Fig. 14c, the elements that are present outside the wear track are compared. As the surface of the disk is coated with the WC–10Co4Cr alloy, whose average microhardness is \(1256 \;{\text{HV}}_{0.02}\), against a sphere that is made of the Cu–35Zn alloy with \(172\; {\text{HV}}_{0.02}\), the sphere wear is abrasive and severe. Even with the high-hardness tungsten carbide, abrasive wear occurred on the disk, thereby demonstrating that the debris that was detached from the disk hard surface was embedded in the surface of the sphere and remained attached to the sphere. Via this mechanism, the sphere behaved as a cutting tool for the disk, and the lubrication film did not prevent the scratching of the surface.

Fig. 14
figure 14

SEM-EDS micrograph and chemical compositions of a WC–CoCr surface that was lubricated with HLP

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A micrograph of the wear track that was produced while using HEPR biodegradable oil is shown in Fig. 15a. The occurrence of abrasion mechanisms is also observed in the magnification of the wear track, similar to those that were observed with HLP oil, but with more grooves and microcuts along the wear direction. In these tests, measurements were also made via EDS (P3.1 and P3.2) to evaluate the mechanisms of wear. In Fig. 15b, 6.5% copper elements (Cu) and 2.43% zinc elements (Zn) were identified; hence, the concentrations of these materials are higher in this case. A decrease in the thickness of the lubricating film enables the surfaces to approach each other, thereby resulting in more severe wear [32]. Abrasive wear occurs when the hard surface (disk) directly cuts off the ductile surface (sphere), which causes grooves and scratches to form on the surface [59]. In this case, an inadequate amount of lubricant and the formation of a lubricant film directly influence the abrasive wear, along with the concentrations and the combination of antifriction, anti-wear and extreme-pressure additives in each oil.

Fig. 15
figure 15

SEM-EDS micrograph and chemical compositions of a WC–CoCr surface that was lubricated with HEPR

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4 Conclusions

From the results that were obtained in the theoretical and experimental studies, the following conclusions are drawn:

  • The comparison of the mixed and elastohydrodynamic lubrication regimes demonstrated similarities in the coefficient of friction between HLP and HEES, which differed substantially from that of HEPR oil. The variations in the boundary lubrication and elastohydrodynamics were smaller in HEPR; hence, it is suitable for applications in the reciprocating sliding process. The smaller variation of the coefficient of friction influences the energy consumption of the hydraulic equipment. Additionally, the starvation phenomenon in the HLP oil was identified from the change in the coefficient of friction with the thickness of the film.

  • The rheological properties of the fluids, such as the pressure-viscosity coefficient, along with the dynamic viscosity, influenced the formation of the lubrication film. The Lambda parameter (λ) did not accord with the values that were obtained in other studies. This difference is due to the high porosity and the surface finish of the WC–CoCr coating, which provides a high value of roughness of the disk, which is 11 times greater than that of the sphere.

  • The concentrations of the Zn and P additives influenced the friction results and the wear mechanisms. Furthermore, a larger proportion of sulfur (S) is used for the HEPR oil. Typically, sulfur is used as an extreme-pressure additive, which also explains the satisfactory performance in short- and long-term sliding tests.

  • In the short-term wear tests, the coefficient of wear K was approximately 3.2 times greater with the HEES oil than with the HLP oil; with HEPR, the coefficient of wear K was approximately 8 times lower than that with HEES.

  • Contamination of the lubricants with copper particles was identified in higher concentrations in the HEES oil. A portion of these particles that detached from the sphere also adhered to the disk track, thereby changing the coefficient of friction and the wear of the surfaces.

  • In the long-term tests, as the number of hours increased, the stability of the coefficient of friction changes. For the HEPR, the instability exhibited the stick–slip phenomenon. For the HEES oil, the mixed lubrication regime changed for the limit lubrication, sometimes reaching the values of the coefficient of dry friction.

  • The long-term tests showed that the wear coefficients differed from those that were obtained in the short-term. The wear coefficient ratio of HEES/HEPR was 7.1, compared to the HEES/HLP ratio of 16.2. The HEPR/HLP ratio was 2.3. The performance of HLP is due to the concentrations of zinc and phosphorus elements in the additives. However, the HEPR oil showed satisfactory performance in the long-term tests.

  • The adhesion phenomenon of the sphere on the disk was more evident in the sliding with the HEES oil. According to the EDS results, 93% of the copper and zinc elements were deposited on the disk surface with the use of the HEES oil, whereas for sliding with the HLP and HEPR oils, 1.46 and 8.93%, respectively, were deposited.

  • Due to the higher concentrations of extreme-pressure and anti-wear additives and the higher value of the coefficient of pressure-viscosity, the HLP oil realized a superior performance. The HLP oil avoided the adhesion of the sphere on the disk and, consequently, realized lower wear rates in the tribological system.

  • At the end of this study, the need was identified to increase the concentrations of additives in biodegradable oils, especially ester-based oils (HEES), to increase their performance and to improve the wear resistance of the surfaces in contact.