V-belt transmission type small spiral oil press

Abstract

The invention relates to the field of oil pressing machinery, in particular to a V-belt transmission type small spiral oil press which comprises a machine frame, a motor installed on the machine frame, a V-belt transmission mechanism, a speed reducing mechanism, a pressing cage and a thread extrusion mechanism installed in the pressing cage. The driving belt wheel is connected with an output shaft of the motor through a flat key, the driven belt wheel is installed at one end of an input shaft of the speed reducing mechanism, the other end of the driven belt wheel is connected with a large gear of the speed reducing mechanism, and the output shaft and the input shaft are installed in a horizontally opposite mode. And the other end is connected with the thread extrusion mechanism through the diaphragm coupling. The oil press is simple in structure, small in occupied space, safe to use and convenient to maintain, the oil pressing efficiency of the oil press is remarkably improved, the oil quality is improved, the added value of products is improved, and more profit margins are brought to production enterprises.

Description

Technical Field

[0001] The invention relates to the field of oil pressing machinery, in particular to a V-belt driven small screw oil press. Background Art

[0002] Screw presses are currently the most common oil extraction equipment. Their main principle is to use mechanical force to compress oil to produce oil. This equipment is generally available in two types: single-screw and twin-screw. They are widely used in the oil and fat processing industry. Single-screw presses are characterized by their simple structure and operating mechanism, low production costs, ability to process a variety of oils and fats, and adaptability to a variety of production conditions. However, single-screw presses still have many problems in long-term operation, such as suboptimal pressing results, poor screw axial thrust, low theoretical compression ratio, and slippage when the oil content is high. Twin-screw extruders, such as the commercially available SYZ twin-screw extruder, SZX12x2 twin-screw extruder, and SYZX340 cold-type twin-screw oil press, while highly efficient and with low residual oil levels, suffer from high rework rates, high energy consumption, substandard oil quality in some cases, and a lack of safety features. Furthermore, these large, high-capacity presses are not suitable for small manufacturers or households due to the large space required and the complexity of operation and maintenance. If the size is reduced directly, it may cause uneven extrusion and reduce the oil yield. Summary of the Invention

[0003] In order to overcome the deficiencies of the prior art, the present invention provides a V-belt driven small screw oil press.

[0004] In order to achieve the above object, the technical solutions specifically adopted by the present invention are as follows:

[0005] A V-belt-driven small screw oil press comprises a frame, a motor mounted on the frame, a V-belt transmission mechanism, a reduction mechanism, a pressing cage, and a screw extrusion mechanism mounted within the pressing cage. The V-belt transmission mechanism includes a driven pulley and a driving pulley. The driving pulley is connected to the motor's output shaft via a flat key. The driven pulley is mounted on one end of the reduction mechanism's input shaft and connected to the reduction mechanism's large gear at the other end. The output shaft and input shaft are mounted horizontally opposite each other. One end of the input shaft is connected to a small gear via a flat key, and the other end is connected to the screw extrusion mechanism via a diaphragm coupling. The driven pulley is a 63mm small pulley, while the driving pulley is a 125mm large pulley. The input shaft is 185mm long and the output shaft is 222mm long.

[0006] Furthermore, the driven pulley and the driving pulley are installed in a vertically aligned manner and are connected and driven by a V-belt, thereby avoiding direct coupling with the drive shaft and protecting the motor.

[0007] Furthermore, the driven pulley and the driving pulley are both three-groove type, which effectively increases the friction, prevents the end screws from locking, and prevents the pulley from falling off.

[0008] Furthermore, the pressing cage is horizontally installed above the frame, with the filter screen side always facing the bottom.

[0009] Furthermore, a funnel-shaped feed port is welded on the upper end of the pressing cage to facilitate the subsequent connection of a valve or a hose to transport the finished oil, and an oil receiving hopper is installed on the bottom end.

[0010] The screw extrusion mechanism includes a pressing screw shaft, an end cone and a shaft mounted at the end of the screw shaft. The end cone is internally threaded, while the shaft is externally threaded. This allows for easy adjustment of the gap between the end cone and the end of the pressing cage, facilitating the extrusion of oil residue. A slag tray is located at the bottom of the end cone, and its internal slope facilitates the oil residue's ability to fall through the tray and slide off the machine surface. The pressing screw is 655 mm long, and its teeth feature a tapered root-circle structure.

[0011] The present invention has the following beneficial effects:

[0012] The invention has a simple structure, occupies a small space, is safe to use, and is easy to maintain. It significantly improves the oil pressing efficiency of the oil press, improves the quality of the oil, increases the added value of the product, and brings more profit space for the production enterprise.

[0013] In terms of environmental benefits, energy-saving design is adopted to reduce energy consumption and waste emissions in the production process, which is beneficial to environmental protection. BRIEF DESCRIPTION OF THE DRAWINGS

[0014] Other features, objects and advantages of the present invention will become more apparent upon reading the detailed description of non-limiting embodiments with reference to the following drawings:

[0015] Figure 1 This is a three-dimensional diagram of a V-belt driven small screw oil press according to an embodiment of the present invention.

[0016] Figure 2 The figure is a front view of a V-belt driven small screw oil press according to an embodiment of the present invention.

[0017] Figure 3 The figure is a top view of a V-belt driven small screw oil press according to an embodiment of the present invention.

[0018] Figure 4 for Figure 3 Enlarged view of the middle Q.

[0019] Figure 5 This is the specific size information of the squeezing screw, unit is mm.

[0020] Figure 6 Select the graph for the power calculation.

[0021] Figure 7 The specific size information diagram of the input shaft is in mm.

[0022] Figure 8 This is the specific size information diagram of the output shaft, unit is mm.

[0023] In the figure: 1-frame; 2-motor; 3-driven pulley; 4-V-belt; 5-driving pulley; 6-diaphragm coupling; 7-speed reduction mechanism; 9-pressing cage; 10-thread extrusion mechanism; 11-feeding port; 12-pressing screw shaft; 13-tail cone; 14-tail cone shaft; 15-slag receiving tray;

[0024] 21-reduction mechanism input shaft; 22-pinion; flat key 23; 24-deep groove ball bearing; 25-reduction box body; 26-sleeve; 27-large gear; 28-reduction mechanism output shaft; 28-reduction mechanism output shaft. DETAILED DESCRIPTION

[0025] The present invention is described in detail below in conjunction with specific embodiments. The following examples will help those skilled in the art to further understand the present invention, but are not intended to limit the present invention in any form. It should be noted that, in the absence of conflict, the embodiments of the present invention and the features in the embodiments can be combined with each other.

[0026] like Figure 1-Figure 4 As shown, a V-belt driven small screw oil press according to an embodiment of the present invention comprises a frame 1, a motor 2, a V-belt transmission mechanism, a speed reduction mechanism 7, a pressing cage 9, and a screw extrusion mechanism 10 installed in the pressing cage 9;

[0027] The motor 2 is installed on the second layer of the frame 1 through the motor seat, and its output shaft is connected to the driving pulley of the V-belt transmission mechanism through a flat key;

[0028] The V-belt drive mechanism, comprising a driven pulley 3, a driving pulley 5, and a V-belt 4, offers advantages such as smooth transmission, low operating noise, simple structure, and low cost. Both the driven pulley 3 and the driving pulley 5 utilize a three-groove design, which effectively increases friction and prevents the end screws from locking and the pulleys from falling off. They are mounted vertically aligned to avoid direct coupling with the drive shaft, thus protecting the motor.

[0029] The deceleration mechanism 7 includes a deceleration mechanism input shaft 21, a pinion 22, a deceleration mechanism output shaft 28, a reduction housing 25, and a large gear 27. The driven pulley 3 is installed at one end of the deceleration mechanism input shaft 21, and the other end of the deceleration mechanism input shaft 21 is connected to the inner wall of the reduction housing 25 through a deep groove ball bearing 24, and is connected to the pinion 22 through a flat key 23; the deceleration mechanism output shaft 28 and the deceleration mechanism input shaft 21 are installed horizontally opposite to each other, and the deceleration mechanism output shaft 28 is connected to the large gear 27 through a flat key 23. The tail end of the deceleration mechanism output shaft 28 passes through the sleeve 26 and is connected to the threaded extrusion mechanism 10 through the diaphragm coupling 6.

[0030] The pressing cage 9 is mounted horizontally on the frame by a bracket, with the filter screen always facing the bottom. A funnel-shaped feed port 11 is welded on the upper end of the pressing cage 9 to facilitate the subsequent connection of a valve or hose to transport the finished oil. An oil receiving hopper 8 is installed at the bottom.

[0031] The threaded extrusion mechanism 10 includes a pressing screw shaft 12, and a tail cone 13 and a tail cone shaft 14 arranged at the tail of the pressing screw shaft. The tail cone 13 is processed with an internal thread, and the tail cone shaft 14 is processed with an external thread. It is connected to the pressing screw shaft 13 through a deep groove ball bearing, so as to facilitate the adjustment of the gap reserved between the tail cone and the end of the pressing cage, and facilitate the extrusion and removal of oil residue. A slag receiving plate 15 is set at the bottom of the tail cone 13, and the internal slope of the slag receiving plate 15 is designed to facilitate the oil residue to fall off the slag receiving plate and slide out of the equipment table.

[0032] When the present invention is used, the oil first enters the feeding device through auxiliary machinery such as an elevator. Driven by the auger blades, it is continuously fed into the pressing cage. As the screw lead decreases and the base diameter increases, the volume decreases, causing a change in pressure. This pressure continuously squeezes the oil. Due to the positive conveying characteristics of the pressing screw, the oil flows out through the gaps in the pressing cage, while the remaining oil residue is pushed forward and pressed until it forms a dense and compact cake that is discharged from the discharge device at the rear of the pressing cage. The oil is then simply filtered through a filter and collected in a container before being sent to the next process.

[0033] The present invention is described in detail below with reference to specific parameter data.

[0034] Transmission ratio calculation

[0035] To prevent foreign matter from entering the oil and causing the main shaft to jam and the motor to burn out, a V-belt is used between the motor and the pressing screw. The belt drive ratio is I=2, and the gear drive reduction ratio is I=5.

[0036] Compression ratio and power calculation

[0037] Considering that peanuts have a high oil content, the total compression ratio is set to 10. The theoretical power consumption formula is:

[0038] N r=(qxnxRp) / 6000 (1)

[0039] For small and medium-sized equipment, using formula (1) we can get: Nr = 4~7kw.

[0040] Take N r =7KW, pressing chamber pressure:

[0041]

[0042] Where β is the coefficient of oil moisture and temperature; w is the moisture content of the pressed material, ranging from 1.1 to 2.8%. When w = 2.8%, β = 0.0045, and e = 2.71 (natural logarithm). Substituting this into formula (2), p ≈ 1.7 MPa.

[0043] Selection of screw size

[0044] The tooth profile of the selected squeezing screw shaft adopts a tapered root circle structure. The selection of this tooth profile is based on the following: the size α of the squeezing screw tooth profile must be controlled within the range of 0° to 30°, and the size of the β angle should be between 15° and 45°, and the maximum value of the β angle must not exceed the limit of 90°. In addition, the degree of the Y angle should be less than 10°. By accurately setting these parameters, the rationality and efficiency of the squeezing screw design are ensured. The specific size information of the squeezing screw is as follows: Figure 5 shown.

[0045] V-belt design

[0046] (1) Diameter design of small and large pulleys

[0047] Use 63mm small pulley, 125mm large pulley, and 320mm center distance.

[0048] Belt length

[0049] Center distance variation range (elastic deformation occurs during belt operation):

[0050]

[0051] Large and small pulley angle:

[0052]

[0053] From formula (4), we can get:

[0054] α>120°, which is consistent.

[0055] Depend on Figure 6 , determine the coefficients of the band:

[0056] (2) Structural design of small and large pulleys

[0057] It is known that d1=63mm, d2=125mm.

[0058] Table 1 Pulley groove dimensions

[0059]

[0060] Note 1) 6min is the recommended minimum wall thickness of the wheel rim.

[0061] The small pulley structure is designed as follows: the wheel groove of the V-belt pulley corresponds to the model of the selected V-belt, see Table 2.

[0062] Table 2 V-belt model correspondence

[0063]

[0064] The surface is closely fitted with the working surface of the pulley groove, and the angle between the working surface of the V-belt pulley groove is made less than 40°.

[0065] When installing a V-belt, it is usually necessary to avoid it extending beyond the outer diameter of the pulley and also to prevent it from directly contacting the bottom of the wheel groove. Therefore, a standard is stipulated that the minimum height from the base diameter of the wheel groove to the outer diameter and bottom of the pulley should be hmin and hmino respectively.

[0066] At the same time, the roughness of the wheel groove working surface is required to meet the standard of Ra 1.6μm or Ra 3.2μm.

[0067] From Table 4.4, the sizes of solid pulleys for large and small pulleys are as follows:

[0068] Small pulley:

[0069] Large pulley:

[0070] Table 4.4V belt type

[0071]

[0072] Gear transmission calculation

[0073] To ensure transmission stability, a helical gear meshing system was selected. Given the low operating speed, a grade 8 precision spur gear (according to GB 10095-88) was used. The pinion gear was constructed from 45-grade steel, tempered and surface-hardened to a hardness of 45 HRC. The large gear was also constructed from 45-grade steel and subjected to the same treatment to a hardness of 45 HRC.

[0074] Pinion σ Hlim1 =1120Mpa,σ Flim1=360Mpa;

[0075] Large gear σ Hlim2 =1120Mpa,σ Flim2 =360Mpa.

[0076] Determine the module Z and calculate the gear module according to the tooth root bending strength:

[0077]

[0078] Where A m is the helical angle coefficient of bending strength; C m is the matching material coefficient; K is the load coefficient, the common value is K=1.2~2; Y FS is the composite tooth form coefficient; z1 is the number of pinion teeth; is the tooth width coefficient; T1 is the pinion torque (N·m); σ FP is the allowable bending stress (Mpa).

[0079] According to the selection of stepper motor, speed and other requirements, the initial gear ratio i=5; the torque transmitted by the small gear is the torque output by the large pulley T=60N·m; the tooth width coefficient is Take values within the range and initially set φ a =0.5; Pairing material coefficient C m =1(steel to steel); composite tooth shape coefficient Y FS =4.3; allowable bending stress σ FP1 =σ Flim1 =360Mpa;σ FP2 =σ Flim2 =360Mpa.

[0080] The modulus is calculated by formula (5):

[0081]

[0082] The modulus can be greater than 3.2, so take 4.

[0083] And z1=10, i=5, so z2=50,

[0084] Center distance:

[0085]

[0086] Pinion pitch circle diameter: d1 = mz = 4 × 10 = 40 mm;

[0087] Large gear pitch circle diameter: d2 = mz2 = 4 × 50 = 200 mm;

[0088] Pinion tooth height: h a1 =h m=1×4=4mm;

[0089] Pinion tooth root height: h f1 =1.2m=1.2×4=4.8mm;

[0090] Gear full tooth height: h1 = h a1 +h f1 =4+4.8=8.8mm;

[0091] Pinion tooth tip diameter: d a1 =d1+2h a1 =40+2×4=44mm;

[0092] Pinion tooth root diameter: d f1 =d1-2h f1 =200-8=192mm;

[0093] Big gear tooth top height: h a2 =h m =1×4=4mm;

[0094] Big gear tooth root height: h f2 =1.2m=1.2×4=4.8mm;

[0095] Gear full tooth height: h2 = h a2 +h f2 =4+4.8=8.8mm;

[0096] Big gear tooth tip diameter: d a2 =d2+2h a2 =40+2×4=48mm;

[0097] Large gear root circle diameter: d f2 =d2-2h f2 =200-2×4.8=191.2mm;

[0098] Gear width:

[0099] Take b2=20mm, b1=50mm.

[0100] Calculate the power transfer capability based on contact strength:

[0101] P HP ≥K A P (6)

[0102] Where, P HP Permissible transfer power; K A is the service factor; P is the transmitted power.

[0103] Since the working characteristics are uniform and stable, the service factor K is taken A =1, the transmission power of the gear P1 = 14 × 10 -5 , so the transmission power of the gear can be obtained from formula (6): K A P = 1 × 14 × 10 -5 =14×10 -5 W.

[0104] Permissible transfer power:

[0105]

[0106] Coefficient C H1 , it is necessary to consider the influence of the pinion speed and the number of teeth ratio on the tooth surface contact strength. This influence can be reflected by the following calculation formula.

[0107]

[0108] From formula (8), we can get C H1 =1.108.

[0109] Coefficient C H2 , it is necessary to consider the characteristics of the node area, the factors of overlap and the effect of the helix angle on the contact strength of the tooth surface. When the helix angle αn is equal to 25°, according to the Mechanical Design Manual, the contact strength coefficient C H2 is 0.21.

[0110] Coefficient C H3 , is to consider the influence of tooth width, small wheel pitch circle diameter, tooth load distribution and inter-tooth load distribution on contact fatigue strength. Its value is calculated by the following formula

[0111]

[0112] Where K Hβ Tooth load distribution coefficient; K Hα The load distribution coefficient between teeth. Take K Hβ =1.36, K Hα =1.2.

[0113] From formula (9), we can get:

[0114]

[0115] Coefficient C H4 , the influence of the lubrication conditions between gears (such as the viscosity of the lubricating oil, the node speed and the roughness of the tooth surface) on the gear load capacity should be considered. It is very important to choose the appropriate lubricating oil viscosity.

[0116] v 50 =100mm2 / s, take C H4 =0.79,

[0117] Take the contact fatigue strength σ Hlim =1120Mpa.

[0118] After in-depth analysis and evaluation of the tooth surface hardening coefficient, it was confirmed that the small hardened gear with smooth surface treatment produced a significant cold work hardening effect on the tooth surface of the large modulated steel gear during operation. This optimization makes the large gear show better stability and durability under load, thereby improving the overall performance and reliability of the entire gear transmission system. w = 1. Size factor Z x , is to consider the influence of factors such as gear size on the contact strength of the tooth surface. Take Zx = 1. Contact strength life coefficient Z NT , is to consider the influence of the gear working life on the allowable contact stress.

[0119] N L =60nat(10)

[0120] Where n is the gear speed; a is the number of contacts between the tooth surfaces on the same side of the gear per revolution; and t is the number of hours the gear is working.

[0121] From formula (10), we can get: N1 = 60 × 300 × 1 × (8 × 300 × 10) = 1 × 10 8 ;

[0122]

[0123] Take the contact fatigue life coefficient Z NT1 =0.98, Z NT2 =1.11.

[0124] Use factor K V

[0125]

[0126] Coefficient K1 = 39.07, K2 = 0.0193;

[0127] Circumferential force

[0128] Circumferential speed

[0129] Set tooth width b1 = 20 mm and use coefficient K A =1, the number of small gears z1 = 10, the gear ratio u = 5 and other values are substituted into formula (11) to obtain K V =1.11.

[0130] Elastic modulus

[0131] Minimum safety factor S for contact fatigue strength Hmin Under normal working conditions and with a failure probability of 1%, the minimum safety factor S Hmin =1.

[0132] Substitute the above values into P HP Calculation formula, get P HP >K A P, the verification result is safe.

[0133] Axis calculation

[0134] (1) Speed, power, and torque of each shaft

[0135] Speed of each axis:

[0136] Input 1 axis: n1 = n 皮带 =300r / min.

[0137] High-speed output shaft 2:

[0138]

[0139] Power of each shaft:

[0140] According to the formula: P 入 η=P 皮带 =300x60 / 9550=2.

[0141] Where, P 入 is the input; η is the transmission efficiency; P 皮带 is the output.

[0142] Input power of each axis, namely:

[0143] 1 axis:

[0144] p1=p θ =2kw

[0145] Input shaft 2: p2=2η=1.9kw.

[0146] Torque of each shaft:

[0147]

[0148] The torque of each shaft can be calculated from formula (12):

[0149] 1 axis:

[0150] Input shaft 2: T2 = 300 N.M.

[0151] (2) Design and verification of input shaft

[0152] According to the parts installed on the shaft, the structure of the shaft is as follows Figure 7 As shown:

[0153] Given an input shaft with a power of 2 kW, a speed of 300 rpm, and a torque of 64 Nm, a safety factor C of 107 was selected after studying a shaft made of 45 quenched and tempered steel (hardness 240 HBS). The minimum shaft diameter can be calculated based on the calculation principles outlined in "Fundamentals of Mechanical Design."

[0154]

[0155] Substituting the data into formula (13) we get:

[0156] d min =20mm.

[0157] Since there is a keyway on the shaft, in order to reduce its influence on the shaft, the d value is increased by 5%, so that d = 20x1.05, that is, 21mm, and finally d = 24.

[0158] Structural design of the shaft:

[0159] To ensure the safe operation of the input shaft, it is necessary to accurately measure the forces acting on the bearings and determine the diameter of their critical sections. For high-speed pinions, the circumferential forces they bear can be expressed by the following formula:

[0160]

[0161] Where d1 represents the pitch circle diameter of the gear, and its value is m millimeters.

[0162] According to the formula, the radial force F is known R The calculation method is: F R =F t *tanα, where F t is the axial force, α is the pressure angle, and its value is 20°. Substituting the given data, we get F t1 =800(N), F r1 The value of is 290 (N). Now we need to solve the support force on the vertical surface.

[0163] By the formula:

[0164]

[0165] Substituting the data into the equation:

[0166] F 1v =145(N)F2v≈140(N),

[0167] Calculate the bending moment about a vertical surface:

[0168]

[0169] Substitute the data into formula (16):

[0170] M av =14.19(Nm)M′av=14.6(Nm),M aH =3.9(Nm), M' aH =3.9 (Nm),

[0171] The formula for the bending moment generated by force F is:

[0172] M=FK(17)

[0173] Substituting the data into formula (17) we get:

[0174] M 2F =24.7 (Nm),

[0175] The bending moment generated by the force F is given by the formula:

[0176]

[0177] Substitute the data into formula (18) to obtain: M aF =8.2(Nm), put M aF and Direct addition, M a =28.12(Nm)M′ a =28.56 (Nm).

[0178] Find the torque transmitted by the shaft:

[0179] The most dangerous section is the one with a keyway hole, where the required bending force is:

[0180]

[0181] Considering the torque stress as a pulsating cyclic stress and converting it into the corresponding compromise value, the following conclusions can be drawn:

[0182] M e =5.5(Nm).

[0183] Calculate the diameter of the shaft at the critical section:

[0184] The shaft made of 45 steel is quenched and tempered. According to the data in the reference, σ B =1150Mpa, (σ-1b)=110Mpa, which can be calculated by the following formula.

[0185]

[0186] Substituting the data into formula (20), we obtain: d ≥ 19 mm.

[0187] According to the weakening of the shaft by the keyway, d is increased by 5%, so d = 1.05 × 19 = 20 mm, where d is the designed minimum shaft diameter greater than d.

[0188] (3) Design calculation of output shaft

[0189] According to the parts installed on the shaft, the structural diagram of the shaft is as follows Figure 8 shown.

[0190] The power P, speed n and torque T on the input shaft are known;

[0191] N=300r / min; T3=64N·mp=2kw.

[0192] Preliminary calculations have roughly determined the minimum shaft diameter. Considering 40Cr (hardness 240 HBS), we've chosen a safety factor, C, of 107, referring to Table 14-2 in "Fundamentals of Mechanical Design." This diameter can be expressed as d = (F / (π*σ))*C, where F is the load on the shaft and σ is the allowable stress of the 40Cr material. This is a theoretical value; actual design should also consider operating conditions and other factors.

[0193]

[0194] Substituting the data into formula (21) we get:

[0195] d min =22 take d2=25mm.

[0196] Calculate the force acting on the dangerous section of the output shaft:

[0197] The circular force is given by the formula:

[0198]

[0199] According to the torque (Nm) and pitch circle diameter (m), it can be calculated

[0200] F t1 =1600(N).

[0201] The radial force of the output shaft is based on the formula:

[0202] F r1 =F T1 tanα (23)

[0203] Where α is the pressure angle, 20°.

[0204] Importing the data into formula (23) yields: F r1 =650(N).

[0205] From these data, the circumferential force F t1 Equal to 1600 Newtons, radial force F r1 Equal to 650 Newtons, and the pitch circle diameter of the large gear is 200 mm.

[0206] Support force of vertical surface: F 1v =300(N)F 2v =300(N).

[0207] By formula:

[0208]

[0209] Substituting the data into formula (24) we get:

[0210] M av =30(Nm)M′ av =30(Nm).

[0211] By the formula:

[0212]

[0213] Substituting the data into formula (25) we get: M aH =90(Nm)M′ aH =90(Nm).

[0214] The bending moment generated by force F is:

[0215] By the formula:

[0216] M 2F =FK (26)

[0217] Substituting the data into formula (26) we get: M 2F =80(Nm).

[0218] The bending moment generated by the force F in section aa is:

[0219] By the formula:

[0220]

[0221] Substituting the data into formula (27) we get: M aF =50(Nm).

[0222] Find the composite bending moment diagram in the worst case, and change M aF Combined with other factors to achieve the best results.

[0223] By the formula:

[0224]

[0225] Substituting the data into formula (28) we get: M a =120(N×m)M′ a =128(N×m).

[0226] Find the torque transmitted by the shaft:

[0227]

[0228] If the torque stress of the shaft is considered to be a pulsating cyclic stress, take the conversion coefficient α = 0.6 and substitute it into formula (29) to obtain: M e =131(Nm).

[0229] In the locations involving dangerous sections, 45Cr that has been quenched and tempered is selected as the material of the shaft to ensure that it has sufficient strength and toughness. B =1150Mpa, σ-1b=110Mpa.

[0230]

[0231] Substituting the data into formula (30), we obtain: d = 22 mm.

[0232] According to the weakening of the keyway, the value of d is increased by 5%, so d = 1.05 × 22 = 24 mm, and the value of d2 is greater than d, thus meeting the standard requirements.

[0233] Bearing calculation

[0234] The bearing mainly plays a supporting role, so it is only subjected to radial force and slight axial force. For the reliability of the design, the axial force is calculated as half of the radial force.

[0235] In terms of bearing selection, it was decided to use deep groove ball bearings. The inner and outer rings and rolling elements are all made of G20Cr2Ni4A (carburized bearing steel). After heat treatment, the hardness is ensured to be no less than 60HRC to meet the requirements of high-load and high-speed working environments.

[0236] For the input and output shafts, we decided to use the same type of bearings, and used the bearings selected for the output shaft as a benchmark for calculation and verification. This ensures that the bearings have the same performance and life during load and rotation.

[0237] According to the previously known radial force F of the large gear r =650N and F a=325N, which can be equivalently converted into the axial and radial loads borne by the bearing, thereby more accurately evaluating the working condition and life of the bearing. This equivalent conversion method makes the selection and calculation of bearings more scientific and accurate.

[0238] Designed lifespan: 8X300X10=24000h.

[0239]

[0240] Where, F a is the axial force; f r is the radial force; e is the judgment coefficient of deep groove ball bearing 0.44.

[0241] Equivalent dynamic load:

[0242] p=f α (XF r +YF a ) (32)

[0243] Substituting the data into formula (32) we get:

[0244] p≈1.2×(0.56×325+1.5×650)≈1298.4N.

[0245] Assuming load factor f q The basic dynamic load rating is a crucial indicator used in equipment design to determine the appropriate range of sustained motion loads that the equipment must withstand.

[0246]

[0247] Substituting the data into formula (33) we get:

[0248] C=5.74KN.

[0249]

[0250] The basic rated dynamic load of the 6000-2RS bearing is 6.2KN, and the basic rated static load of the 6000-2RS bearing is 2.9KN. C<[C0] can be met.

[0251] Verification: relative axial load corresponding to

[0252]

[0253] Equivalent dynamic load p0 = 1.2 × (0.56 × 325 + 1.5 × 650) = 1298.4 N.

[0254] Life calculation:

[0255] So it can be applied.

[0256] The above shows and describes the basic principles, main features, and advantages of the present invention. Those skilled in the art should understand that the present invention is not limited to the above embodiments. The above embodiments and descriptions are merely preferred examples of the present invention and are not intended to limit the present invention. Various changes and improvements may be made to the present invention without departing from the spirit and scope of the present invention. Such changes and improvements fall within the scope of the present invention. The scope of protection claimed in the present invention is defined by the appended claims and their equivalents.

Claims

1. A V-belt driven small screw oil press, characterized by: It includes a frame, a motor installed on the frame, a V-belt transmission mechanism, a reduction mechanism, a pressing cage and a threaded extrusion mechanism installed in the pressing cage. The V-belt transmission mechanism includes a driven pulley and a driving pulley. The driving pulley is connected to the output shaft of the motor through a flat key. The driven pulley is installed at one end of the input shaft of the reduction mechanism, and the other end is connected to the large gear of the reduction mechanism. The output shaft and the input shaft are installed in a horizontally opposed manner. One end of the input shaft is connected to the pinion with a flat key, and the other end is connected to the threaded extrusion mechanism through a diaphragm coupling.

2. A V-belt driven small screw oil press according to claim 1, characterized in that: The driven pulley and the driving pulley are installed in a vertically aligned manner and are connected for transmission via a V-belt.

3. A V-belt driven small screw oil press according to claim 1, characterized in that: The driven pulley and the driving pulley are both three-groove type.

4. A V-belt driven small screw oil press according to claim 1, characterized in that: The pressing cage is horizontally installed above the frame, with the filter screen side always facing the bottom.

5. A V-belt driven small screw oil press according to claim 1, characterized in that: A funnel-shaped feeding port is welded on the upper end of the pressing cage, and an oil receiving hopper is installed on the bottom end.

6. A V-belt driven small screw oil press according to claim 1, characterized in that: The thread extrusion mechanism includes a squeezing screw shaft, a tail cone and a tail cone shaft arranged at the tail of the squeezing screw shaft, an internal thread is processed in the tail cone, an external thread is processed in the tail cone shaft, a slag receiving plate is arranged at the bottom of the tail cone, and a slope is designed inside the slag receiving plate.

7. A V-belt driven small screw oil press according to claim 1, characterized in that: The driven pulley is a 63mm small pulley, the driving pulley is a 125mm large pulley, the input shaft is 185mm, and the output shaft is 222mm.

8. A V-belt driven small screw oil press according to claim 6, characterized in that: The squeezing screw is 655 mm long and has a conical root circle structure.

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