1. Introduction
A high water-based hydraulic system (HWBHS) has better fire-resistance, lower environmental impact, lower operating cost, greater rigidity, and more accurate positioning ability compared with oil hydraulic systems. Since the 1980s, HWBHS has been widely used in food processing, ocean engineering, and mining engineering [
1,
2,
3,
4].
As the key actuating component, HWBHM has been studied for use in HWBHS [
5,
6,
7,
8,
9,
10,
11,
12]. An HWBHM with low speed and large torque has the advantages of compact structure and convenient layout. It has special benefits that can be used in limited spaces where low speed and high torque output is required but there no space for a reduction device. In contrast, HWBHM need reduction devices to get low speed and high torque. HWBHM with low speed and high torque has wide application prospects, e.g., in tunnel boring machines, hydraulic winches, and hand-held emulsion drilling machines, etc. In traditional axial/radial pump/motors working with high water-based hydraulic liquid (HWBHL), the flow distribution mechanism is usually a valve plate, and most HWBHMs operate at high speed due to the characteristics of the valve plate. To improve the volume efficiency and abrasive resistance of friction pairs in an HWBHM, the key wear and lubrication problems of the valve plate in a high water-based hydraulic motor/pump have been studied. The structure of the valve plate has significant influence on the volumetric efficiency of a traditional hydraulic pump (Marning [
13], Seeniraj [
14], Wang [
15], and Yang [
16]), and numerical simulation of the relationship between leakage flow and clearance in flow distribution pair was implemented by Wang [
17]. In addition, a satellite motor that can be fed with various liquids was designed by Śliwiński, and the flow of liquids (including oil and emulsion) in flat gaps was analysed [
18]. The low viscosity and poor lubrication characteristics of HWBHL directly limited the development of HWBHMs, especially HWBHMs working with low speed and high torque. Anyhow, leakage caused by the valve plate in a HWBHM is inevitable. Thus, to promote the volume efficiency and working life of HWBHM working at low speed (less than 100 rpm), an HWBHM with a self-balanced valve (SDV) distribution mechanism has been introduced in our previous work [
19]. For the SDV distribution mechanism, a cone valve has strong self-compensation capability and is leak-proof. Thus, the leakage caused by the valve plate in HWBHM could be alleviated.
In addition, to improve the working efficiency and working life of HWBHM, the friction and leakage problems of a piston-cylinder pair with lubrication of HWBHL still need to be addressed. As shown by Wang, based on hydraulic oil film theory, with the same clearance, the leakage caused by a piston-cylinder pair is larger than other friction pairs such as the distribution pair and slipper pair in construction machinery remanufactured piston motors [
20]. Analogue experiments involving the piston-cylinder pair in internal a curved radial piston motor were implemented by Olsson [
21]: the experimental results showed that the effects of abrasion would be greater when the motor is working with water-glycol at low speed. An optimal clearance formula of a piston friction pair for water hydraulic pump based on thermal balance principle was proposed by Huang [
22,
23]. The analysis showed that a minimum thermal clearance exists in each piston friction pair, and the factors of friction coefficient, matching materials and piston diameter, etc., should be considered. Cao [
24] studied the stress state and leakage in a piston-cylinder pair in hydraulic axial piston pump, pointing out that proper clearance and materials should be considered to solve the friction performance in a water hydraulic pump. Experimental research on materials used for both a piston and cylinder in a water hydraulic pump has been implemented by Yang [
25]. The results of this study indicated that it is more suitable to use stainless steel matched with plastics as the materials of piston and cylinder in a water hydraulic axial piston pump. Several materials (metal, engineering ceramics, and plastics) and engineering technologies for the main friction pairs in a water hydraulic piston pump were tested: the result showed that a hard-to-soft scheme forms a feasible matching pair with water lubrication [
26,
27,
28]. To improve the load bearing of, and reduce leakage from a piston-cylinder pair in axial piston machine at high pressure with water as the hydraulic liquid, Ernst [
29,
30] investigated the influences of different surface shapes on pressure build-up, leakage, and torque loss. The previous research laid a foundation for the further study of hydraulic elements with lubrication of HWBHL.
To deal with the friction and leakage problems, and study the influences of various factors (seal length, clearance, temperature and materials) on leakage performance of PSC pair in HWBHM with SDV, it is necessary to provide guidance regarding the design of and materials chosen for the PSC pair in HWBHM with SDV. Firstly, the basic structure and motion performance of a PSC pair were analysed, and the structure of the PSC was designed. Then, a friction experiment was applied to analyse the frictional characteristics of the matching pairs, namely (1) a hard-to-soft matching scheme of 316L stainless steel with OVINO-GIC coating (316L-GIC) and PEEK reinforced with 30% carbon fibre (PEEK-30CF), (2) a hard-to-hard matching scheme of 316L-GIC and 316L-GIC. Finally, the influences of various factors (clearance, temperature, pressure, and materials) on the leakage performance of PSC pair in HWBHM were analysed based on an orthogonal test design considering fluid-structure interaction, and the feasibility of clearance sealing for PSC pair in HWNHM working at low speed was studied.
2. Model of PSC Pair in HWBHM with SDV
The proposed new type HWBHM with SDV [
18] is shown in
Figure 1. The HWBHM mainly consists of two parts: the SDV groups are shown in
Figure 1b and the piston pairs are shown in
Figure 1c. As shown in
Figure 1, there are five pistons in the HWBHM, and the distribution of each piston is realised through one inlet distribution valve (I-DV) and one outlet distribution valve (O-DV). The I-DV and O-DV have the same structures, the DVs all consist of five parts: a valve guiding, spring, valve, valve seat, and rod. When I-DV and O-DV are in on or off states, the hydraulic force upon the valve
Fp1 is equal to the hydraulic force under the valve
Fp2. The support force from valve seat
FN0 is equal to the spring force
Ft in an off state. The motion of the valve in DV is controlled by a rod, while the force exerted by the rod is decided by the spring force and hydrodynamic force, and the spring force and hydrodynamic force are very small at low working speeds. For the applied valve seal, the proposed SDV could reduce the significant volume efficiency loss problem caused by the valve plate distribution mechanism. Besides, unlike the case in an axial piston motor, the lateral force for PSC pair in HWBHM is relatively small. The wear seen in PSC pairs could be alleviated, but leakage and friction between piston and swivelling cylinder still need to be addressed.
The conflicts between lubrication and wear, sealing and leakage are common challenges in the engineering of friction pairs. For a PSC pair in HWBHM with SDVs, there are mainly two kinds of sealing mode: non-contact sealing (a clearance seal) and contact sealing mode (using a sealing ring). For a clearance seal, the friction pairs are under lubricated conditions most of the time, which has the advantage of minor abrasion between friction pair components. Thus, the working life of a friction pair with clearance seal mode would be longer than that in contact seal mode (the clearance would, however, affect the volume efficiency of the HWBHM).
The basic structure of a PSC pair is shown in
Figure 2a: it consists of a piston and a swivelling cylinder. The diameter of the piston is
d, the seal length in the PSC pair is δ
lp, and the clearance between piston and swivelling cylinder is
hp.
The motion of a piston in its cylinder at an arbitrary position is shown in
Figure 2b,
O represents the centre of rotation of the crankshaft,
O1 represents the eccentricity of the structure on the crankshaft,
O2 represents the centre of rotation of the cylinder,
e0 represents the distance between
O and
O1,
R0 is the distance between
O and
O2,
θi represents the angle of rotation of the cylinder,
φi denotes the angle of rotation of the crankshaft,
ω represents the angular velocity of the crankshaft, and
li is the distance between
O1 and
O2. A PSC pair at position
φi = 0 is shown in
Figure 2c. For the proposed HWBHM with SDV, the designed displacement
q is 189 mL/r, the rated rotation speed is 60 rpm, and maximum rotation speed is 100 rpm. The corresponding basic parameters in
Figure 2 are listed in
Table 1.
As shown in
Figure 2b, when the crankshaft is at position
φi, the relative velocity
vi between the piston and swivelling cylinder can be expressed as [
18]:
In Equation (1), li is calculated as: .
Thus, the variation of velocity
vi with angular position
φi can be obtained (
Figure 3). The maximum relative velocity
vi between piston and swivelling cylinder will be less than 0.2 m/s when the working speed of HWBHM is no more than 100 rpm.
4. Friction Experiment for Matching Materials
Since the viscosity of HWBHL is very low, friction pairs in a motor/pump combination always work under insufficient lubrication, which would cause abrasion that directly affects the operating life of the HWBHM. In the case of lateral force caused by other friction pairs such as the friction force in a piston slipper and crankshaft pair, abrasion could occur under HWBHL lubrication conditions when the piston and swivelling cylinder make contact. Thus, the selection of matching materials is an important part of their engineering design and specification.
As shown by the friction test of different matching materials on the pin-disk friction machine (
Figure 12) in reference [
34], for a test speed of
nt =15 rpm (corresponding to a linear velocity of 0.0314 m/s) with load
P = 200 N (corresponding to a contact pressure of about 21 MPa for the friction pair), 316L-GIC (OVINO-graphite intercalated compound) /PEEK-30CF (polyetheretherketone reinforced with 30 per cent carbon fiber) has the optimal friction characteristics. This is followed by matching materials 316L-GIC/316L-GIC. The abrasion would be severe if both matching materials were made of stainless steel. According to the relative motion between the piston and cylinder as shown in
Figure 3, friction coefficients at 15 rpm and 90 rpm (corresponding to a linear velocity of 0.1884 m/s) were analysed to reveal the friction performance of the two matching materials within the rotation scope of the HWBHM-SDV. The pin specimen and disk specimen for test are shown in
Figure 13.
The friction test for each matching group will last for 10 h, while the abrasion and friction coefficient for each matching pair would be observed every 2 h to check the abrasion surface of specimens after friction. Thus, the specimens will be taken from the test machine, and then cleaned through an ultrasound cleaner and dried. During the test, specimens will be taken from the test machine every 2 h and then reinstalled on the test machine for the next test period. Thus, the reinstallation error would inevitably influence the surface contact of matching materials, which could also cause the discontinuity of CoF (coefficient of friction) between the end of the previous test period and the beginning of the subsequent test period. In this study, the influence of reinstallation error on matching pairs would not be considered.
The friction coefficients of matching materials 316L-GIC and PEEK-30CF under different test speeds are shown in
Figure 14. At 15 rpm (
Figure 14a), the friction coefficient decreases with the increase of test time, and the friction coefficient reaches a stable value of 0.01 at 6 h to 8 h. At 8 h to 10 h, the friction coefficient is about 0.03. At 90 rpm (
Figure 14b), the friction coefficient is smaller than that at 15 rpm, and the friction coefficient of 316L-GIC and PEEK-30CF is about 0.02 after an abrasion time of 6 h. The sharp increase in friction coefficient is caused by the centripetal effect, because of the insufficient lubrication of liquid at the touch surface between pin and disk specimen. It could be concluded that the increase of speed of relative motion between the friction pairs is conducive to enhanced lubrication, and could decrease the friction coefficient.
The friction coefficients of matching materials 316L-GIC and 316L-GIC under different test speeds are shown in
Figure 15. At 15 rpm (
Figure 15a), the friction coefficient of 316L-GIC is stable throughout the test, and the friction coefficient is about 0.065 after 10 h abrasion. At 90 rpm (
Figure 15b), the friction coefficient is lowest from 0 h to 2 h, and then reaches a stable value of about 0.06. The sudden change in friction coefficient from 2 h to 4 h and 8 h to 10 h is mainly caused by the abnormal stopping and starting of the abrasion machine during the test.
Due to the running-in between the pin specimen and disk specimen being conducive to reducing the surface roughness of PEEK-30CF, the friction coefficient of hard-to-soft matching materials 316L-GIC and PEEK-30CF decreases over time at test speeds 15 rpm and 90 rpm, while for hard-to-hard matching materials 316L-GIC/316L-GIC, the friction coefficient is more stable during the 10 h abrasion test. The increased speed would increase the lubrication between the pin and disk specimens.
5. Influence of Factors on Leakage Performance of PSC Pair in HWBHM with SDV Based on Orthogonal Test
The analysis of friction coefficients for the two group matching materials shows that they all have good friction performance within the relative motion speed of the piston and swing cylinder. In addition to the abrasion performance, leakage performance is another important indicator. Especially under HWBHL lubrication, the leakage in friction pairs is always a severe problem. Many factors (such as structure parameters, working pressure, temperature, material properties, etc.) would affect the leakage in the PSC pair. There might exist complex interact effect among the factors [
35,
36]. To analysis the coupling influences of factors (including clearance, temperature, and working pressure) on leakage performance for the two kinds of matching materials as shown in
Figure 16, the bidirectional fluid-structure coupling analysis is applied. Because the influence of working speed is very small, it will not be considered in this study.
To investigate the influence of fluid-structure interaction on leakage in a PSC pair, the method of fluid-structure coupling analysis was implemented through commercial software ANSYS 15.0. The simulation was conducted by the combination of a Fluent module and Transient Structure module. There are two pairs of coupling surfaces: one is the outside coupling surface on the lubrication film and coupling part on the cylinder, the other is inside the coupling surface on the lubrication film and coupling part on the piston. After repeated coupling calculations, the leakage δqhpc at the pressure outlet port for the lubrication film was recorded.
An orthogonal test was used to study the effects of the main influencing factors. The result can also provide guidance on the structural design of the PSC pair: based on the above analysis, clearance, temperature, and working pressure were selected as three test factors, and because each factor contains three levels, a type L
9 (3
3) orthogonal test table was established (
Table 3). The value of leakage in the clearance was calculated as shown in
Table 4, where δ
qp is the leakage from a PSC pair with matching materials 316L-GIC/peek-30CF, and δ
qs is the leakage from a PSC pair with matching materials 316L-GIC/316L-GIC.
To determine the importance of factors affecting leakage, the averages of factors at each level and the ranges are shown in
Table 5 where
xqpi represents the mean of δ
qp for each factor at level
i (
i = 1, 2, 3), and
Rqp represents the range of
xqpi for each factor with matching materials 316L-GIC/PEEK-30CF.
xqsi denotes the mean of δ
qs for each factor at level
i (
i = 1, 2, 3), and
Rqs represents the range of
xqsi for each factor with matching materials 316L-GIC/316L-GIC. The effect of factors will increase with the increase of
Rqp or
Rqs.
Thus, under the analysis of bi-direction fluid-structure coupling analysis, the degree of influence on
Rqp(s) is: A > C > B for the matching materials 316L-GIC/316L-GIC and 316L-GIC/PEEK-30CF. To reflect the influence of each factor intuitively, the effects of all factors on means of δ
qp(s) for different matching materials are described in
Figure 17. It can be concluded that the influence difference of clearance on leakage is not significant between matching materials 316L-GIC/316L-GIC and 316L-GIC/PEEK-30CF at Levels 1 and 2. The influence of pressure increases with the increase of clearance, causing larger deformation of materials with an increase of pressure. For the temperature, the difference between the two matching pairs is significant at low levels (
T = 10 and
T = 20), but the difference is not significant at Level 3 given the influence of pressure distribution under fluid-structure coupling analysis. For the influence of pressure, the difference between the two matching pairs increases with the increase of pressure.
During the operation of an HWBHM, there are about 2.5 PSC pairs under high pressure conditions, the theoretical volume efficiency loss
ηhpv caused by PSC pairs can be expressed by Equation (6):
To compare the difference in volume efficiency loss between the calculation considering fluid-structure interaction analysis and theoretical analysis shown by Equation (7), parameter
kct was established, where,
ηhpvc is the volume efficiency loss caused by leakage δ
qhpc considering fluid-structure interaction:
Considering the influences of surface roughness and machining error, the clearance is set to be 10 μm in this analysis. With the assumption that the temperature of HWBHL is set to be constant at 20 °C, the deformation of structure caused by temperature is ignored.
When considering the matching materials 316L-GIC and PEEK-30CF, the variation of volume efficiency loss and difference between fluid-structure interaction and theoretical analysis is shown in
Figure 18. For matching materials 316L-GIC/PEEK-30CF, the volume efficiency loss could be very high when the working pressure exceeds 15 MPa, and the difference increases with the increase of working speed at working pressures exceeding 10 MPa and working speeds exceeding 30 rpm. In addition, the volume efficiency loss would increase by about 450 % at 100 rpm and a working pressure of 30 MPa, thus, limited by the material properties of matching materials, the maximum working pressure should be less than 15 MPa when the working speed is less than 100 rpm.
For matching materials 316L-GIC and 316L-GIC, the variation of volume efficiency loss, and difference between fluid-structure interaction and theoretical analysis are shown in
Figure 19. For the small deformation of stainless steel, it can be seen that the working pressure should be limited to less than 28 MPa at a working speed of 100 rpm to obtain a high volume efficiency, while the working pressure should be limited to less than 20 MPa at a working speed of 40 rpm; however, the feasible scopes of working pressure and working speed are much wider than that of matching materials 316L-GIC and PEEK-30CF. Parameter
kct increases with the increase of working pressure and working speed. The variation of
kct is more stable under lower working speeds and lower working pressures. The volume efficiency loss will increase by more than 120% at a working pressure of 30 MPa and a working speed of 100 rpm.
To improve the working performance of PSC pair in the HWBHM-SDV, and improve the working scope of matching materials, an automatic compensation structure could be considered to compensate the clearance caused by structure deformation.