THE EFFECT OF AXIAL CLEARANCE IN GEAR PUMPS ON VOLUMETRIC EFFICIENCY Milutin ŽIVKOVIĆ 1, Jasmina MILJOJKOVIĆ 2, Tijana STOŽINIĆ 3, Ivan BIJELIĆ 3, Slobodan MILOŠEVIĆ 3, Nemanja MOR 4 1 High Technical Mechanical School of Professional StudiesTrstenik, milutinzivkovicts@gmail.com 2 University of Kragujevac jasmina.miljojkovic@fink.rs 3 High Technical School of Professional Studies Novi Beograd, tijana.stozinic@visokatehnicka.edu.rs, 3 High Technical School of Professional Studies Novi Beograd, ivan.bijelic@visokatehnicka.edu.rs, 3 High Technical School of Professional Studies Novi Beograd, slobodan.milosevic@visokatehnicka.edu.rs 4 High Technical School Zemun, nemanja.mor@visokatehnicka.edu.rs Abstract Experimental testing and exploitation of gear pumps show that the increase of pressure reduces the volumetric flow. Therefore, the company Prva Petoletka (Trstenik, Serbia) invests continuously in constructive improvements of the gear pumps. Its subsidiary, Hidraulika (Trstenik, Serbia), has been producing gear pumps for more than 60 years. This paper analyses the impact of the axial clearance in external gear pumps, by defining the equation for the actual flow rate as a function of pressure. Conducted experimental research should support the mathematical analysis, and point to further constructive and productive improvements, i.e. solutions that will produce maximum volumetric efficiency. Keywords gear pumps, volumetric efficiency, axial clearance, actual flow A I. INTRODUCTION LL hydraulic pumps convert mechanical energy of the drive motor into hydraulic energy of the working fluid, in other words, produce flow and pressure. According to the way in which the hydraulic energy is produced, pumps may be classified as kinetic (dynamic) pumps and constant-volume pumps (positive displacement pumps). However, in systems for general use, only the positive displacement pumps are commonly employed, M. Živković D. Todorović, and S. Vasić [1]. The principle of their function is based on the working chambers volume variation, which enables the suction and suppression of the working fluid, with forced movement of the operating element. The forms of operating elements can be: gears, bolts, pistons, or vanes. Gear pumps are characterized by a simple design, high operating reliability, and low production costs, as compared to other types of positive displacement pumps. They are produced with both low and high operating pressure, but higher levels of pressure increase flow losses and cause the increase in the working fluid 13 temperature on the outlet. High pressures in bearings and the substantial flow pulsation induce the fluctuations in pressure, causing the noise that eventually reaches a substantially higher level. In low-pressure conditions, gear pumps are used in hydraulic systems for lubrication of tribological assemblies of working machines, as well as in systems for oil and petroleum transfusion. The actual flow in gear pumps is lower than the theoretical flow by the amount of the working fluid that flows from the area of high pressure to the low-pressure area, due to the structural clearance between the gear teeth front surface and the casing. This paper analyses the effect of axial clearance on volumetric efficiency of external gears pumps, produced by PPT-Hidraulika AD Trstenik. Data were collected via experimental measurement. The results should point to an optimal value of the axial clearance, given that modern gear pumps, in terms of attained pressure, almost take on the characteristics of piston pumps, according to S. Kolarević [2]. II. GEAR PUMPS PRODUCED BY PPT HIDRAULIKA AD (TRSTENIK) In the production of gear pumps, a commonly used constructive solutions are combined with the application of empirically based improvements, Fig. 1, 2, 3 and 4. The most commonly used are the involute spur gears, where standard sizes correspond with standardised parameters range, according to T. Xianzhao [3]; Fig. 1.Cross-section of the external gear pump Specific working volume: 0,25 266,66 (cm 3 /o)
(for special purposes it can be higher); Operating pressure: 20 200 bar, at present 270 bar (development of the solutions for p = 320 bar is in progress); Rotational speed of the drive shaft: 500 5.000 (min -1 ) (depending on nominal size and place of installation); Total efficiency: 0,90 0,99; Noise level 87 (db). "PPT Hidraulika", Trstenik, uses the configuration with sealing, which restrict the compensation surface housed in the bearing bushing. Fig. 2. Example of a compensation surface housed in a bearing bushing This solution has significant advantages, because the engraving of the cover is avoided, which allows the selection of materials with better mechanical properties, thus enabling the achievement of higher pressures. With bedding in covers, it is possible to achieve smaller dimensions, and, therefore, to provide considerably more stable operation of the pump, and consequently more stable operation of the system as well, Fig.3. It is all enabled by the installation of roller bearings, which are suitable to be built in mobile machines. The bearings role is to compensate the external influences (vibrations, impacts). to N. Manring, S. Kasaragadda [4]. Fig. 4. External gear pumps with a pressure valve. The level of nominal pressure was raised to 250 300 bars, with a tendency to further rise up to 400 bars. However, the high pressure makes the noise problem even more acute; therefore, the application of these pumps is narrowed to the fields where the power is the primary factor, and the noise is less relevant. It is primarily the application in hard operating conditions, typical for hydraulic systems in mobile machines. The last twenty years of development of these pumps in PPT Hidraulika AD has resulted in the creation of a new family of gear pumps, labelled as ZPB, Fig. 5, as in PPT Hydraulic AD [5]. This family of gear pumps can be applied in hydraulic systems which require pressures up to 300 bars. Fig. 5. Gear pump from the ZPB family Fig. 3. Example of a compensation surface in bearing bushing, with bedding in covers All manufacturers constantly strive to increase the pressure level, while simultaneously retaining other parameters, such as noise, service life, and overall efficiency, at the current level or higher. Success in these efforts largely depends on the level of technological development, on which every innovation in the field of construction must rely. This primarily refers to: development of new high-quality materials for vital elements, new processing procedures, and the advance in the methods and techniques of measurement. External gear pumps, due to their simple design, have a chance to retain the primacy as the least expensive pumps, despite the structural-technological requirements. This primarily relates to raise of the pressure until it reaches the level which is typical for piston pumps; thereby, the field of the gear pumps application could be expanded, according 14 These pressures, until recently reserved exclusively for piston pumps, are achieved through the new concept of the casing and the pump cover. Casing material is extruded aluminium with high deformation degree (extruder). By using the material with improved mechanical characteristics, the achieved permanent dynamic strength is significantly higher than in aluminium die-casting. Thus, both the pump and the device in which the pump is installed, attain higher reliability. Hydraulic casting with E-module of (1416) 10 7 (N/m 2 ) is used for the cover. When compared to the aluminium covers with E-module of (79) 10 7 (N/m 2 ), the new concept improves the reliability of sealing at high pressures, due to a higher values of E-module. Furthermore, the use of hydraulic casting for front and rear covers is more favourable for the threads of the fixing screws, which improves the compactness of structure. In the rear cover, made of hydraulic casting, the valves, such as flow control valves or pressure limiting valves, can be fitted, as shown in Fig. 5 (significant in the pumps for servo control installations of mobile machines). It is well known that the number of gear teeth has the greatest effect on the flow pulsation. In this family of pumps, the gears with 12 teeth are adopted, and therefore, the pulsations of flow and pressure are
reduced (12%). For comparison purposes, as shown below, the flow pulsations are paired with corresponding numbers of teeth (the data relate to previous constructive solutions): 1) 9 teeth, pulsation is 22%, 2) 10 teeth, pulsation is 18%, 3) 11 teeth, pulsation is 16%, 4) 12 teeth, pulsation is 14%. Considering the decisive effect of the pulsation on noise, the presented concept gives a major contribution to the reduction of noise. The bearing journals of gears have a specified maximum roughness below 1 (μm), which enables the optimal performances in sliding bearings coated with Teflon (DU bearings). The uniform surface hardness of gears is achieved by quality heat treatment, which in the end results in high resistance to wear of sleeve, bearings and gearing. DU bearing is a bronze bearing coated with Teflon with steel support. The technique with DU bearing was applied for gear positioning because it guarantee the optimal exploitation. DU bearing is a bronze bearing coated with Teflon with steel support. This technique of positioning has significant advantages, N. Manring, S. Kasaragadda [4]: low friction coefficient, good sliding properties, good capability of installing, slight wear, high dynamic loading, broader temperature range, moisture resistance, long service life. In this generation of pumps, pulsations of pressure and flow are reduced, and consequently, the reduction of noise was also achieved. That was enhanced by the solution for bearing bushings with the optimal pressure distribution and by the adequate design of drainage of the "squashed" oil on pressure side. Therefore, a considerable improvement is achieved as compared to the technical solutions, which have a drain of "squashed" oil toward the suction side. The characteristic of bearing bushings in these pumps is also a partially bevelled edge on the side next to the gear, which provides a controlled increase of pressure in a gear tooth space. A new generation of gear pumps (ZPB) has anti-cavitations characteristics, Fig. 5: 1) Easy creation of pressure in the teeth chambers; reduced sensitivity to the influence of air; 2) High hardness of bushings prevents the erosion in the mesh area, and thus sensitivity to dirt is decreased; 3) Specially shaped distribution edges on bearing bushings, enables the "opening" teeth chambers to fill at the very beginning, i.e. prematurely; therefore, cavitations at high rotation speed are avoided. III. FLOW LOSSES IN EXTERNAL GEAR PUMPS Flow losses in gear pumps occur as the consequence of clearance (axial and radial) between the casing and the rotating parts. Successfully solved construction of these pumps, which achieve maximum volumetric efficiency (which greatly depends on losses caused by the axial 15 clearance), are the result of years of testing and constant improvements, along with the experience of personnel engaged in production process and the level of equipment in the machinery park. Mathematical analysis and experimental tests show that there are a number of factors that affect, to a greater or lesser extent, the losses caused by axial clearances. Essentially, those factors are: pressure, viscosity, the width of the sealing zone, the rapidity of changes in clearance, gear rotation, etc. In this paper, the focus will be on the axial clearance, which affects the losses more than any other factor. But we can influence on the axial clearance in a great extent. Since the losses increase as the cube of the axial clearance, when the axial clearance decrease, the losses will decrease as well. Axial clearance can be reduced significantly by enhancing the accuracy of parts manufacturing, along with continuous improvement of the processed surface's quality. It must not be allowed that the deviations of roundness, bore-to-face perpendicularity, or parallelism, lead to a gear jam inside the pump casing (due to the small clearance). The lateral shift of the gear, increased by the height of unevenness (which results from the ovality), must be less than the value of axial clearance, so that the gear could rotate freely inside the pump casing. The clearance should not be reduced below the size of the oil layer on the lateral surfaces. Insufficient lubrication results in a semi-dry friction and partly even in a dry friction, wherein a great amount of heat develops. One part of the heat transfers to the gear, while the other part transfers to the wear plates, causing them to expand; thus, the clearance is even more reduced and it will certainly lead to a failure of the pump. Temperature conditions, in which the pump is operating, must be taken into account to. This is particularly important in pumps that work within the systems exposed to high temperatures (200 280). The thermal expansion, corresponding to a maximum operating temperature, cannot be reduced more than by the size of the anticipated clearance, i.e. the size which should ensure effective lubrication of the moving elements. Mechanical deformation of the elements under the action of hydraulic forces must also be taken into account. That action is particularly noticeable in pumps that operate at high pressures, according to T. Xianzhao [3]. Under the action of hydraulic forces, the wear plates tend to separate from the gears, thus increasing the clearances, and consequently, the leakage. The increase in the clearances is amplified by additional forces caused by additional pressure created in the annular zone of the lateral gear side. That pressure, caused by flow pulsation, can be extensive, and therefore the forces caused by him can be substantial, all depending on the size of the lateral surface. This is one more factor that should be taken into account when determining the minimum size of the axial clearance. It should be noted that during the pump exploitation, due to the wear of moving parts, between which the clearance
increases, losses begin to increase to. The longer the pump works, the higher the losses are. The clearance should be decreased by the size of the wear, in order to achieve its desired size during the pump exploitation. In order to minimize the flow losses in gear pumps, primarily by reducing the leakage through the axial clearances, other methods are applied as well. Those methods are based on the application of new solutions, especially in pumps where the hydraulic compensation of the axial clearance is possible, as well as in pumps with high-pressure sealing. One of the most commonly used methods for increasing the volumetric efficiency is the hydraulic compensation of axial clearance. It directly affects the axial clearance, which stays nearly constant during the entire operation of the pump. The constant here means that it does not change under the influence of the wear of rubbing surfaces or the changes caused by a thermal expansion. Essentially, these pumps are structurally the same as the pumps without hydraulic compensation of axial clearance. The difference is that they have mobile wear plates, which will here be referred to as bushings. The pressure of the fluid coming from suppression line of the pump and going behind the bushings presses the bushings against the lateral side of the gear. It is enough to create the pressure on one side of the bushing, and therefore it is avoided to bring fluid on both sides, which significantly complicates the construction. The base problem in these types of pumps comes to the selection of optimal nominal size of the surface, by which the force that presses the bushings against the lateral side of the gear is determined as well. Dimensioning this area is very difficult, because it is not possible to accurately determine the force that separates the plates, which is caused by the pressure of the fluid located in the axial clearance. The force (F) that presses the bushings against the gears should be slightly stronger than the separating force, so that the bushings would constantly be in contact with the gear, achieving the necessary sealing with the least possible losses. The proper selection of the surface (A) which will be exposed to the pressure is very important, because if the surface is not big enough, the sealing force will be insufficient and the fluid will easily leak through the axial clearance. If the area is oversized, the stronger force F will squeeze the oil (between the leading bushing and the gear) and create conditions for semi-dry and dry friction. The relation between these forces should be selected in such a way that neither will come to the separation of the bushing from the gear nor the jamming will occur at the sliding surface of the bushing. The force of the gear-plate separation depends on the pressure on the lateral side of the gear, and its distribution. The force depends on the pressure on: the periphery of the gear, the radial clearance, the eccentricity of the gear, viscosity, etc. Therefore, it is impossible to determine the exact force that separates the gear from the bushing. In the gear tooth spaces, in the pump suction area, due to the rotation of 16 the gear, the sub-pressure is created and that space fills with working fluid through the suction line. Thus, the working fluid, caught between the teeth of the gear set, is transmitted to the discharge side. Theoretical flow of gear pumps is defined by the following expression (1): Q T = V n = q p 2π n = q p ω (1) where: Q T theoretical pump flow; V (cm3/o) displaced volume; q p specific pump flow equal to the shaft volume that is displaced while the shaft rotates by one turn (m 3 /rad); n revolution number of the pump shaft (s -1 ); ω angular velocity of the pump shaft (rad/s). Displaced volume depends on the geometry of gear tooth space; therefore, several different empirical formulas can be used for the approximate flow calculations of gear pumps with the same gears. The actual pump flow is obtained when the theoretical pump flow is reduced by flow losses, the Expression (2): Q S = Q T ΔQ (2) where Q S, Q T, and ΔQ are actual flow, theoretical flow and flow loss [m 3 / s]. Flow losses occur as the result of fluid flow from the area of higher pressure to the area of lower pressure through the structural clearances between the front surfaces of teeth and the casing, sides of the teeth and the casing, as well as in the area of teeth contact. Volumetric efficiency of the pump (η v ) is defined as the relation between the actual and the theoretical flow, Expression (3): η v = Q S = Q T ΔQ = 1 ΔQ (3) Q T Q T Q T Flow losses depend on: the clearances, the deference between pressures in suction and suppression side of the pump, and the working fluid viscosity. Studies show that the Poiseuille law can be applied to flow losses in all hydraulic pumps, which means that flow losses are proportional to the difference between pump pressure and geometrical characteristics, and inversely proportional to the fluid viscosity. The actual pump flow can be defined by the expression (4): Q S = Q T ΔQ = Q T C p η (p p p u ) η v = Q T C where: C geometrical characteristic of the pump (m 3 ), Δp pressure difference in the pump (Pa), p p pressure at the discharge side of the pump (Pa), η v volumetric efficiency. If it is assumed that physical characteristics of the oil are constant and if the pressure at the suction side of the pump (which is insignificant when compared to the pressure on the discharge side), (4)
than from the expression (4) the actual flow can be obtained, Expression (5): Q S = Q T C p p (5) where: C geometrical characteristic of the pump [m 3 ], p p pressure at the discharge side of the pump (Pa). TABLE I PUMP NO. 1 AXIAL CLEARANCE OF 110 μm 0 31,9 0,99 50 31,4 0,975 100 31,1 0,966 150 31,2 0,969 210 31,5 0,98 250 31,6 0,981 270 31,5 0,98 n = 1150 min -1 t = 50 0 C Theoretical pump flow: Q T = 28 [cm 3 /o] 1150[min -1 ] = 32,2 (l/min) TABLE II PUMP NO. 1 AXIAL CLEARANCE OF 110 μm 0 52,5 0,987 50 52,7 0,99 100 52,5 0,991 150 52,6 0,99 210 52,8 0,99 250 52,7 0,99 270 52.6 0,99 Theoretical pump flow: Q T = 28 [cm 3 /o] 1900[min -1 ] = 53,2 (l/min) TABLE III PUMP NO. 2 AXIAL CLEARANCE OF 130 μm 0 32,0 0,994 50 31,4 0,975 100 30,9 0,96 150 30,9 0,96 210 31,3 0,972 250 31,7 0,984 270 31,8 0,988 n = 1150 min -1 t = 50 0 C Theoretical pump flow Q T = 28 [cm 3 /o] 1150[min -1 ] = 32,2 (l/min) TABLE IV PUMP NO. 2 AXIAL CLEARANCE OF 130μm 0 53,0 0,996 50 52,6 0,99 100 52,2 0,98 150 52,1 0,98 210 52,6 0,99 250 52,7 0,99 270 52,6 0.99 Theoretical pump flow Q T = 28 [cm 3 /o] 1900[min -1 ] = 53,2 (l/min) TABLE V PUMP NO. 3 AXIAL CLEARANCE OF 140μm 0 32,0 0,994 50 31,0 0,963 100 30,4 0,944 150 30,3 0,94 210 31,0 0,963 250 31,4 0,975 270 31,4 0,975 n = 1150 min -1 t = 50 0 C Theoretical pump flow Q T = 28 [cm 3 /o] 1150[min -1 ] = 32,3 (l/min) TABLE VI PUMP NO. 3 AXIAL CLEARANCE OF 140μm 0 52,8 0,992 50 52,3 0,983 100 51,6 0,97 150 51,6 0,97 210 52,0 0,98 250 52,3 0,983 270 52,0 0,98 Theoretical pump flow Q T = 28 [cm 3 /o] 1900[min -1 ] = 53.2 (l/min) TABLE VII PUMP NO. 4 AXIAL CLEARANCE OF 160μm 0 31,9 0,99 50 31,2 0,99 100 30,6 0,95 150 30,7 0,953 210 31,2 0,99 250 31,5 0,99 270 31.5 0,99 n = 1150 min -1 t = 50 0 C Theoretical pump flow Q T = 28 [cm 3 /o] 1150[min -1 ] = 32,3 (l/min) TABLE VIII PUMP NO. 4 AXIAL CLEARANCE OF 160μm 0 52,8 0,994 50 52,2 0,98 100 51,8 0,974 150 51,8 0,974 210 52,3 0,983 250 52,0 0,978 270 52,1 0,98 Theoretical pump flow Q T = 28 [cm 3 /o] 1900[min -1 ] = 53,2 (l/min) It can be concluded that the value of the actual flow of the pump depends on pump's constant geometrical characteristic and on the value of pressure at the discharge side. Results of the measurements of the actual flow in the pump with structural working volume of 28 (cm 3 /o), at various values of pressure at the outlet (from 0 to 27 bar) are shown in Tables I to VIII. Measurements were conducted according to technical conditions for the testing of these types of pumps. 17
By using the data from Tables I to VIII, Fig.s 6 to 9 show the changes in actual flow as the function of pressure (for the drive shaft's constant number of revolutions per minute). The same data are given in comparative diagrams in Fig.s 10 to 11. Fig. 11. Comparative diagram of the flow loss (for n =1900 min -1 ) Fig. 6. Actual flow of the Pump 1 (as the function of pressure) Fig. 7. Actual flow of the Pump 2 (as the function of pressure) Fig. 8. Actual flow of the Pump 3 (as the function of pressure) Fig. 9. Actual flow of the Pump 4 (as the function of pressure) Fig. 10. Comparative diagram of the flow loss (for n = 1150 min -1 ) IV. CONCLUSION The results of the measurements of the actual flow in external gear pumps show that the increase in pressure at the discharge side of the pump cause the increase in flow loss. Inevitably, it leads to the decrease in actual flow, and consequently, to the decrease in volumetric efficiency. It can be said that, hydraulic pumps manufactured in production facilities of the company Hidraulika, Trstenik, due to its characteristics, do not fall behind the global competition. The presented analysis shows that with the axial clearance of 110 μm the most satisfactory volumetric efficiency is achieved, all by utilising the existing production facilities. Literature, as well as years of experience of the authors of this paper, show that in future we can expect further development of gear pumps, oriented towards the slight increase in pressure, but also focused on retaining the high level of volumetric efficiency ( v 0,95), further reducing the pulsations of pressure and flow, and prolonging the service life. These performances can be achieved by working on: modification of gearing, improvement of the construction of bearing bushings, introduction of new technological processes, application of new materials. Finally, it can be concluded that business success of the companies that manufacture hydraulic and pneumatic devices and systems is based on selection of proven design solutions and testing their quality, as well as the use of materials with enhanced characteristics, modern technological equipment and processes, the use of modern measuring and control equipment, and the utilization of modern devices for both static and dynamic programme testing, etc. 12. REFERENCES [1] M. Živković, D. Todorović, and S. Vasić, Gear pumps in the past and present, Journal EMIT, Vol. 3, No.4, pp. 187-195. ISSN 2217-9011. [2] S. Kolarević, Analysis of solution the axial clearance compensation in gear pumps, Conference HIPET '89, Vrnjačka Banja, 13-15 April, 2000. pp. 33-44. (in Serbian). [3] T. Xianzhao, Modeling and Simulation of Gear Pumps based on Modelica/MWorks, Proceedings 8th Modelica Conference, Dresden, Germany, March 20 22, 2011, pp. 421 429. [4] N. Manring, and S. Kasaragadda, The Theoretical Flow Ripple of an External Gear Pump, Journal of Dynamic Systems, Measurement and Control Transaction of the ASME, September, 2003. pp. 396-404. [5] PPT Hydraulic AD: Commercial Technical Documentation (Gear Pumps 2015), (in Serbian). 18