1. Introduction
Single screw compressors (SSCs) are widely applied in air compression, refrigeration, petroleum and petrochemical industries due to advantages such as high efficiency, excellent mechanical balance and compact structure. A SSC is mainly constituted of a screw, a pair of gate-rotors and a casing, as shown in
Figure 1. A screw and a gate-rotor are named a meshing pair. The gas is compressed in the working chamber formed by the screw groove, gate-rotor tooth and the inner wall of the casing. The working principle is shown in
Figure 2. The inlet gas (blue) fills the working chamber as the tooth closes the screw groove. With further rotation, the working chamber shrinks and the gas is compressed (pink). As the working chamber connects the discharge orifice on the casing, the compressed gas (red) is discharged. Liquid is injected into working chambers of SSCs to cool the gas, lubricate the meshing pair and seal the gaps.
A main drawback of the SSC is the deterioration of its discharge capacity. A lot of SSCs sold in the market showed a sharp decrease of discharge capacity for more than 10% after one or two years of initial running, due to the abrasion of the gate-rotor and the enlargement of the meshing pair clearance even under liquid injection [
1]. A lot of worn gate-rotors were collected [
2] and one of them is shown in
Figure 3.
A volume of research has been conducted on improving working life of the SSCs. Structure improvement such as floating gate-rotor was proposed to increase the flexibility of the gate-rotor and reduce wear [
3]. The introduction of the wear-resistant material PEEK to the gate-rotor and the improvement of the machining precision raised the stability of SSCs to some extent [
2]. However, these efforts have been proved to have limited contribution. Several researchers have suggested that meshing pair profile (MPP) is the key factor of the abrasion problems [
4]. Zimmern proposed the original MPPs including single straight line type (SSLT) and single column type (SCT) [
5]. The SSLT is shown in
Figure 4a; the meshing surface is only a straight line. Therefore, the meshing area approaches zero and the meshing surface curvature along the rotating axis of the gate-rotor is infinite, which cause quick abrasion according to the tribology principle [
6]. As shown in
Figure 4b, the meshing line of SCT is not fixed and can slide onto the triangular column surface. Feng developed a multi straight lines type (MSLT) MPP, which makes the meshing line switch between all the straight lines [
7], as shown in
Figure 4c. Generally speaking, the increase of the straight line number will bring better dry friction resistance performance. Combining the advantage of the SCT and the MSLT, a multi-column type (MCT) MPP was proposed [
8], as shown in
Figure 4d. The contact area is maximized and the curvature is minimized. This study proposed the concept of the MCT and its basic design method, without further properties investigations. Li designed a test rig with eccentric wheel to study the wear resistance property of different MPPs [
9]. The range of the enveloping angle, contact stress and relative velocity were accorded with the real condition of SSCs. After 20 h running, the wear loss of sample teeth with different profiles was tested by three coordinate measuring machines. The measurement results show that MCT profile has the best wear resistance and the SSLT profile has the worst performance. Wang proposed a theoretical method to predict the wear characteristics of the MPPs by calculating friction angle and Hertz contact stress [
10].
Along with the development of MPPs, much research has been devoted to the lubrication characteristics. The lubrication characteristics are even more important than the wear resistance since the gate-rotor tooth will never contact with the screw thread flank if the lubrication is well formed between the meshing pair. Reference [
11] discussed the possibility of hydrodynamic lubrication in the clearance between the screw groove flank and the gate-rotor tooth flank, pointing out that the necessary conditions for hydrodynamic film including wedges, sufficient sliding velocity and liquid flow from big inlet to small outlet are available. Heidrich suggested that for the water flooded SSCs, the tribology property may get worse because of low viscosity of water and an acceptable working life may not available [
12]. Jin calculated the oil film thickness under a certain load by Martin equation and verified the existence of the oil film between the meshing pair by an experiment utilizing the electrical insulation of the oil [
13]. Post and Zwaans researched the hydrodynamic properties in SSCs, calculating the oil film pressure distribution by finite difference method and comparing the hydrodynamic lubricating characteristics of SSLT and SCT [
14,
15]. Sun investigated the oil film force at both sides of the tooth and indicated that the oil film force on leading flank is always smaller than on trailing flank [
16]. This numerical result accords with the actual phenomenon that the gate-rotor was always worn seriously on the leading flank [
17]. Wu studied the Couette-Poiseuille flow in two-dimensional asymmetric gaps and gave the approximate solution of the pressure distribution, which could serve as a reference for investigating the pressure distribution in meshing pair gaps [
18]. Huang optimized the SCT MPP and developed an oil flooded prototype [
19]; nevertheless, this study only concentrated on one MPP and no comparison study was conducted. Li designed a modeling experiment to simulate the motion of the meshing pair with water lubrication and verified that water in the gap can establish hydrodynamic lubrication and tested the water film force [
2], but this experiment ignored the Poiseuille effect and the characteristic of synchronous meshing of the three teeth was not taken into consideration. Further, the hydrodynamic lubrication properties of the SSLT and MCT profiles in water flooded SSCs were compared. The numerical results show that the MCT profile generates greater water film pressure and thrust than SSLT profile [
20]. However, this work did not propose a comprehensive and intuitive method to evaluate the lubrication properties of the MPPs. In addition, the influences of the working condition and the machine size were not investigated.
The wear resistance performance of the MPP under dry friction, boundary or mixed lubrication can be judged by testing the wear loss of gate-rotors or calculating the Hertz contact stress. However, how does one evaluate the fluid film lubrication properties of the MPPs? For one gap between the friction pair such as a journal bearing, the lubrication properties can be well described by the liquid film pressure, thrust or thickness. Nevertheless, there are two gaps at leading and trailing sides of the tooth (See
Figure 5). The liquid film pressure, thrust or thickness of one gap cannot represent the tooth lubrication properties. Actually, there are always three teeth of the gate-rotor meshing with the screw at the same time (See
Figure 2) and six gaps are generated. Therefore, it is more complex for evaluating the MPPs’ lubrication properties.
It can be inferred that the gate-rotor will rotate slightly in the screw grooves if the liquid in the six gaps applies a resultant torque on the gate-rotor and the gate-rotor will deflect on the axis of the gate-rotor. Taking one of the teeth to study, it deflects about the gate-rotor axis slightly from EFGH to E’F’G’H’ (See
Figure 5). Consequently, a geometric parameter named micro deflecting angle
δ is proposed to describe the micro deflecting motion. By calculating the micro deflecting angle
δ of every moment in a period, the micro deflecting motion trajectory (MDMT) of the gate-rotor is available. The MDMT of the gate-rotor is similar to the journal bearing center track to some extent. For problems such as deflection trajectory or center track, some scholars have conducted relative research. Jiang and Pi established a model of the tool tip ellipse trajectory deflection control, measured it and analyzed the relation between tool tip ellipse trajectory deflection and the cutting quality [
21]. Xie defined a new concept of instantaneous whirling speed of axis orbit and studied its new perspective for the vibration analysis of cracked rotors [
22]. To bore elliptical hole, Liang and Lu applied Gauss pseudospectral method to obtain the relation between load capacity and servo system, then made the shaft center orbit quickly get close to the designed elliptical hole [
23]. To improve the stability of the journal bearing in twin-screw compressors, Wang presented a homogeneous two-phase flow model to calculate bearing axis orbit and analyzed the impact from evaporating temperature and different built-in volume ratio [
24]. By analyzing the MDMT, the working state of the gate-rotor and the tribology information can be obtained. The lubrication characteristics of the MPPs in SSCs can be evaluated by the MDMT comprehensively.
In this paper, a mathematical model of the MDMT is established. Based on the model, numerical calculations have been carried out to evaluate the lubrication performance of different MPPs in machines with different sizes or under different working conditions.
3. Results and Discussion
In this section, the lubrication performance of different MPPs in machines with different sizes or under different working conditions is presented. Since the MSLT is a transitional MPP and has not been developed for industrial applications, the SSLT, SCT and MCT are set as the study objects. Machines of three discharge capacities are designed and listed in
Table 2.
3.1. Pressure Distribution of Different MPPs
In the 6 m
3·min
−1 SSC, the water film pressure distributions at leading flank under rated discharge pressure were calculated under
δ = 0° and
φgr = 14.9°. Under this gate-rotor rotating angle, the property that the instantaneous meshing line cross different columns of MCT can be well demonstrated and the pressure of the compression chamber is 325300Pa. As shown in
Figure 11, the calculating data are plotted to a 3D surface for every MPP. The 3D surface is projected to the Y
0OZ
0 plane and a 2D pressure distribution of the tooth flank is available.
For SSLT, the peak value of the water film pressure is 332300Pa and emerges at tooth tip. Although the relative velocity gets its minimum at tooth tip, the included angle of the wedge reaches its maximum at the same location. Along with the increases of l, the hydrodynamic effect declines gradually until it vanishes at l = 22.3mm. As l > 36.1 mm, the tooth is out of the screw and there is no liquid film. Since the meshing line is constant, the water film in the 2D plane is a thin and long rectangle. Due to its weak hydrodynamic effect and small action area, the thrust force to the leading flank is small.
For SCT, the peak value is 362000Pa and also emerges at tooth tip. The acting area of the water film presents a triangle in the 2D plane since the meshing point varies from bottom to top of the tooth in the tooth thickness direction (th) along with the increment of l. Compared to the SSLT, the hydrodynamic effect is much stronger and the acting area is bigger.
For MCT, the peak value is 431102Pa and still emerges at tooth tip. As l < 25.8mm, the meshing line is on the column near bottom of the tooth and the action area is very wide. As l > 25.8mm, the meshing line is on the column near top of the tooth and the action area is narrow. Under this φgr, the MCT has the strongest hydrodynamic effect and the biggest action area among the 3 MPPs.
It can be found that the hydrodynamic effect declines from tooth tip to tooth root for all types of MPP. It is mainly because both the contraction ratio of the wedge and the relative velocity decrease continuously with the increment of l.
3.2. Lubrication Performance under Different Discharge Pressure
The lubrication performance for different MPPs adopted in the 6 m3·min−1 SSC under discharge pressure of 0.8 MPa, 1.2 MPa and 1.6 MPa (absolute pressure) is investigated in this section.
3.2.1. Total Torque Tt in a Whole Period
The included angle between two adjacent teeth is
γ. Calculating the lubrication properties, the calculation period is not 2π but
γ, since the gate-rotor coincides to itself after a rotation angle of
γ. In this calculation,
δ is set 0 throughout. The calculation result is shown in
Figure 12.
Total torque Tt is negative in the whole period, which shows that water films at trailing sides are more powerful than those at leading sides without deflecting (δ = 0). That Tt is negative also implies that the water film at leading side will bear the load of |Tt|. The smaller value of Tt means |Tt| is bigger and more difficult to bear by the leading side films.
It is easily found that Tt decreases with the discharge pressure raising for any definite MPP. This proves that Tb grows faster than Ta and the load |Tt| for the leading side films raises with the discharge pressure increment. Taking MCT as an example, the peak of |Tt| increases by 494.52% from 0.8MPa to 1.6 MPa.
It can also be found that under any discharge pressure, the SCT has the minimum Tt. The SSLT has the maximum of Tt under 1.6 MPa and 1.2MPa and the MCT has the maximum of Tt under 0.8MPa. Under the discharge pressure of 1.6MPa, Tt of SSLT varies from −2.905 N·m to −0.477 N·m, Tt of SCT varies from −8.526 N·m to −2.258 N·m, Tt of MCT varies from −5.506 N·m to −0.531 N·m. The SSLT has the minimum load |Tt| and the SCT has the maximum load |Tt|. However, SSLT cannot be regarded as the best profile since it is still related to the water film stiffness.
In addition, all the
Tt curves in
Figure 12 are negative. This can be explained by the fact that although there are three teeth meshing with the screw, the contributions of different tooth to
Tt is not the same. The first tooth (See
Figure 6) is the dominant one due to the high gas pressure applied on it. In the first tooth, the meshing line at trailing side is longer than that at leading side. Therefore, the water film torque generated at trailing side of the first tooth is the maximum one as
δ = 0. This is also the basic reason why all the
Tt curves are negative.
3.2.2. Water Film Stiffness
Total torque
Tt is calculated in the range of −
δlim to
δlim under the gate-rotor rotating angle
φgr0 when one of the teeth just closes the screw groove. The numerical results are shown in
Figure 13.
Tt increases with the increase of δ, because the water films at leading side get thinner and generate bigger torque Ta while the films at trailing side become thicker and generate less |Tb|. As δ approaches δlim = 0.02°, Tt rises rapidly to prevent contact at leading side. When δ approaches −δlim = −0.02°, Tt falls rapidly to prevent contact at trailing side. If Tt > 0 under δlim and Tt < 0 under under −δlim, the gate-rotor is viewed to have self-regulating capability.
It can be seen that the variation ranges of Tt under the discharge pressure 0.8MPa are −1.773~0.028N·m, −6.03~−0.681N·m and −3.997~3.136N·m for SSLT, SCT and MCT, respectively. Only SCT cannot get a positive value even when δ equals δlim, which means wear occurs in the SCT meshing pair.
When the discharge pressure increases to 1.2MPa and 1.6MPa, all the Tt curves drop and only the MCT keeps the Tt positive under δlim, which represents that the MCT may have the best self-regulating ability under φgr0.
By derivation the
Tt curves in
Figure 13, the curves of the water film stiffness
S are obtained as shown in
Figure 14. The stiffness of the water film in the meshing pair can be understood as follows: the water film in the infinitesimal
dl can be seen as a micro spring, the micro springs at different tooth length
l in a tooth flank constitute a spring combination. All the spring combinations at different tooth flanks form a spring system. The water film stiffness can be regarded as the stiffness of the spring system.
It can be observed that the water film stiffness firstly decreases and then increases as
δ increases (See
Figure 14). In the range of -0.01–0.01°, the water film stiffness almost remains a constant (
Figure 13). The water film stiffness
S at
δ = −
δlim (trailing side) is often larger than that at
δ = +
δlim (leading side).
The water film stiffness rises when discharge pressure increases. Taking MCT as an example, the S values under δ = −δlim are 391.01 N·m·deg−1, 477.57 N·m·deg−1 and 515.09 N·m·deg−1 under 0.8 MPa, 1.2 MPa and 1.6 MPa.
Since contact often occurs at leading side, the water film stiffness at +δlim is very important, which represents ability of resistance to wear. It is found that the MCT has much higher water film stiffness at +δlim when compared to the SSLT or SCT. Under discharge pressure of 0.8 MPa, the S values under δ = +δlim are 157.69 N·m·deg−1, 391.00 N·m·deg−1 and 433 N·m·deg−1 for SSLT, SCT and MCT, respectively.
The results shown in
Figure 13 and
Figure 14 are only under a certain gate-rotor rotating angle
φgr0; nevertheless, the evaluation of the MPP needs to be proceed in a whole period. In addition, calculations for loads and water film stiffness are indirect for hydrodynamic properties evaluation. Therefore, the MDMT evaluation method is proposed for this purpose.
The gate-rotor tooth or the three teeth together can be viewed as a combined double slider. It can bear load in both directions and can even be unloaded. The gaps value is coupled with the load generated by water films and they are influenced by each other.
3.2.3. MDMT Calculation Results
The calculation results of MDMT under different discharge pressure are shown in
Figure 15. Since
δbal in this calculation is always greater than 0, the coordinate value on y axis is set from 0 to +
δlim. If
Tt remains negative under the condition that
δ = +
δlim; it indicates that wear occurs at leading side and the contact force will be involved to balance the negative
Tt. In this case, the
δbal is set +
δlim.
Under the discharge pressure of 0.8 MPa, the MDMT curve of SCT always coincides with the straight line of δbal = +δlim in a whole period γ. This result suggests that the water films at leading side cannot afford sufficient bearing capacity to bear the torque generated by water films of trailing side even when δ = +δlim and contact force from the leading side participates in balancing the negative Tt. The MDMT curve of SSLT coincides with the straight line of δbal = +δlim in a whole period except a small region which φgr ranges from −2.49° to 1.51°. In this small region, the hydrodynamic lubrication is established for the leading side and Tt=0 or δbal is available. However, in this region, δbal is too small; in other words, the water films at leading side are too thin when the gate-rotor gets balanced. Contact is still prone to occur under shock. The MDMT curve of MCT is always kept in the region of 0~+δlim, which represents the MCT has good self-regulating ability and contact will not happen in both sides in a whole period γ. It suggests that the gate-rotor is floating in the water films of both sides and meshing with the screw. In addition, the MDMT curve of MCT is relatively far away from the line of δbal = +δlim, which proves that water films at leading side still have enough thickness to remain at full fluid hydrodynamic lubrication and to resist some impact.
Under the discharge pressure of 1.2MPa, only in a very small region about 3.4° (−2.55°~0.82°) the MDMT curve of SSLT does not coincide with the straight line of δbal = +δlim. The MDMT curve of SCT is still the straight line of δbal = +δlim, which is consistent with the situation under 0.8MPa. The MDMT curve of MCT remains in the region of 0~+δlim, but the curve gets closer to the line of δbal = +δlim than that under 0.8MPa. This suggests the MCT still has good self-regulating ability to prevent contact. However, as δbal is achieved, the water films at leading sides get thinner than those under 0.8MPa.
Under the discharge pressure of 1.6MPa, both MDMT curves of SCT and SSLT completely coincide with the straight line of δbal = +δlim. This illustrates contact occurs at leading sides throughout. For the MCT, it cannot keep full fluid lubrication in the whole period but only in a 17.71° range including the region of −27.2°~−22.49° and the region of −8.49°~4.51°. It can be inferred that all the MDMT curves move towards the straight line of δbal = +δlim when the discharge pressure increases. Compared to the SSLT and SCT, the MCT can keep full fluid lubrication in a wider pressure range and has the best self-regulating ability.
It can be easily observed that the working state of the gate-rotor and the contact information are clearly shown in the MDMT curves. Therefore, the MDMT is an intuitive method to evaluate the lubrication properties of the MPPs.
According to the analysis above, it is almost impossible for the SSLT and SCT to avoid contact and wear. These two profiles are widely adopted in the market sold SSCs, especially the SSLT. The analysis may theoretically reveal the reason why traditional SSCs in the market gain a reputation of low wear resistance and fast discharge capacity decrease.
In addition, the discharge pressure should be restricted by the lubrication performance since the gate-rotor will be easily worn out under excessively high pressure ratio.
3.3. Lubrication Performance in SSCs of Different Machine Sizes
A major impact of the machine discharge capacity on the lubrication performance is that different machine sizes bring different relative velocities. For the 3/6/12 m3·min−1 SSCs, the relative velocities vr at middle of the tooth (midpoint of the tooth tip and root) reach 26.3/31.4/40.6 m·s−1 respectively. In this section, the discharge pressure is set to 0.8MPa.
3.3.1. Total Torque Tt in a Whole Period
In this calculation,
δ is set 0 throughout. The calculation result is shown in
Figure 16.
It can be observed that total torque Tt are always negative for SSLT and SCT. For MCT, Tt are negative in 3 and 6 m3·min−1 machines. However, in the 12 m3·min−1 machine, Tt presents positive in the whole period.
It can be easily found that for any definite MPP, total torque Tt increases with the machine size in the studied range. Taking the SCT as an example, the peak value of Tt increases from −1.93 N·m to −1.13 N·m when the discharge capacity raises from 3 m3·min−1 to 12 m3·min−1. It proves that Ta increases more quickly than Tb as the machine size raises. For MCT, the Tt curve rises much more rapidly than the other two MPPs as the machine size is enlarged. This is mainly because the MCT has the largest wedge at leading side.
It can also be found that in the machines of the same discharge capacity, the MCT often has the maximum Tt and the SCT has the minimum Tt. In the 12 m3·min−1 machines, the peak value of −1.13 N·m, −0.195 N·m and 4.37 N·m for the SCT, SSLT and MCT, respectively. In MCT, Tt is positive and the load Tt will be borne by the water films at trailing sides. Although 4.37 N·m> |−1.13| N·m, the water film stiffness of the trailing side is often higher than that in the leading side. Therefore, it is still difficult to judge which MPP has the best performance.
3.3.2. Water Film Stiffness
Total torque
Tt is calculated in the range of −
δlim to
δlim under the gate-rotor rotating angle
φgr0 when one of the teeth just closes the screw groove. The numerical results are shown in
Figure 17.
It is found that
Tt increases with
δ increasing, which is consistent with that found in
Figure 12. It can also be observed that the machine of big discharge capacity has the larger variation range of
Tt. Taking MCT for example, the variation range of
Tt are −3.22~0.82 N·m, −3.99~3.14 N·m, −4.26~13.41 N·m for 3 m
3·min
−1, 6 m
3·min
−1 and 12 m
3·min
−1 machines, respectively. It implies that for the machine with big size, the self-regulating system is easier to be constituted. It is mainly due to the higher relative velocity and larger hydrodynamic water film in the SSC of big size.
The SCT can establish the self-regulating system only in 12 m
3·min
−1 SSC. The SSLT can build the self-regulating system in both 6 m
3·min
−1 and 12 m
3·min
−1 SSC. The MCT can form this in all the three SSCs. By derivation of the
Tt curves in
Figure 17, the curves of the water film stiffness
S are obtained and shown in
Figure 18.
It is found that the water film stiffness firstly decreases and then increases as δ increases. The curve shape helps to prevent contact since the stiffness is big at both ends but small in the middle. The water film stiffness S at δ=−δlim (trailing side) is often larger than that at δ = +δlim (leading side). This is consistent with the phenomenon that wear at leading side is more serious than that at trailing side.
For any MPP, the water film stiffness increases with the discharge capacity increasing. This implies that the relative velocity has significant impact on the water film stiffness. The water film stiffness is not in a linear relation with the discharge capacity. Taking the MCT as an example, the water film stiffness S(δ=δlim) of the 6 m3·min−1 and 12 m3·min−1 machines raises by 160.84% and 695.01% compared to the 3 m3·min−1 machine.
3.3.3. MDMT Calculation Results
The calculation results of MDMT in machines of different discharge capacity are shown in
Figure 19.
In the 3 m3·min−1 SSC, it is found that the MDMT curves of the SSLT and the SCT coincide with the straight line of δ = δlim in a whole period of γ. It implies that |Tb| is always larger than Ta and contact at leading flank is inevitable. It is observed that the MDMT curve of MCT coincides with the straight line of δ = δlim in a small region about 5° (φgr varies from −19.19° to −14.22°). In the remaining region of the whole period, the MDMT curve of MCT does not coincide with δ = δlim anymore. This indicates that wear occurs in part of the period and the water film at leading flank is unable to bear the load of |Tb| in the whole period.
The calculation results of the 6 m
3·min
−1 SSC under 0.8MPa are shown in
Figure 15 and analyzed in
Section 3.2.3. In order to facilitate the comparison among the three machines, the calculation results are replotted in
Figure 19. It is worth mentioning that in Ref. [
20], a 6 m
3·min
−1 prototype adopting the MCT has been made and a 2000 h endurance test under 0.8MPa has been carried out. In the experiment, the test result does not show any sign of the discharge capacity loss during the whole 2000 h and the tooth flank basically remains in its original shape. The calculated MDMT curve of MCT for the 6 m
3·min
−1 machine in
Figure 19 does not contact with the leading side or with the trailing side in the whole period. The calculated curve is in good agreement with the experimental results.
In the 12 m3·min−1 SSC, it is found that the MDMT curve of the SSLT coincides with the straight line of δ=δlim in a small region about 8.67° (φgr varies from −20.17° to −11.5°). In this range, Ta cannot balance Tb and contact occurs at leading side. The MDMT curve of the SCT reaches δ=δlim when φgr ranges from −18.79° to −13.48°. In the rest range of 27.41°, the SCT MDMT curve is between δlim and -δlim. The MCT MDMT curve does not reach δ = δlim or δ = −δlim in a whole period. It indicates that contact will occur neither at the leading flank nor at the trailing flank. The water films at both sides regulate the gate-rotor effectively. The gate-rotor floats in the water films of both sides.
Overall, in the studied range, with the increment of compressor discharge capacity, the MDMT curves of different MPPs move towards the straight line of δ = −δlim. The liquid films get thicker at the leading side and thinner at the trailing side when δ = δbal. The lubrication performance is improved. The phenomenon that wear occurs at leading flank will reduce. The SSLT and SCT cannot achieve meshing without contact in all of the three SSCs, despite the improvements in the larger machines. The MCT gate-rotor can achieve this goal in the 6 m3·min−1 and 12 m3·min−1 machines.
Further, with increment of the relative velocity, the hydrodynamic effect of leading side increases quicker than that at trailing side. This may mainly be because the relative velocity at tooth tip at leading side is larger than that at trailing side.
3.4. The Fluid Film Lubrication Ratio in a Whole Period
ψ is defined as the ratio of the gate-rotor rotating range in which the meshing pair is lubricated by fluid film to the whole period
γ. Based on the calculation results above, the relations between
ψ and other lubrication parameters are shown in
Figure 20. The influences of discharge pressure and relative velocity
vr (machine size) are analyzed, respectively.
In
Figure 20a, the 6 m
3·min
−1 machine is chosen to analyze the impact of the discharge pressure on
ψ. For SSLT,
ψ equal to 12.2%, 10.3% and 0 under 0.8MPa, 1.2MPa and 1.6MPa. For SCT,
ψ is always a constant of 0. For MCT,
ψ equal to 100%, 100% and 54.11% under the three discharge pressure levels. It can be easily observed that along with the discharge pressure increases,
ψ decreases for the SSLT and MCT. This indicates that the fluid film lubrication range can be shortened by the increasing pressure and machines with excessive high pressure ratio may not achieve a satisfied life.
In
Figure 20b, the discharge pressure is set to 0.8MPa to compare the influence of the relative velocity
vr on
ψ. For SSLT,
ψ equal to 0, 12.22% and 73.51% under 26.3 m·s
−1, 31.4 m·s
−1 and 40.6 m·s
−1. For SCT,
ψ varies from 0 to 83.77%. For MCT,
ψ increases from 84.81% to 100%. In the studied range,
ψ increases with the relative velocity
vr. This implies that the fluid film lubrication range can be extended by increasing
vr, which can be achieved by the larger machine size or frequency conversion.