Design method of liftable steering drive axle of amphibious vehicle

Abstract

The invention discloses a design method of a liftable steering drive axle of an amphibious vehicle, which comprises the following steps: S1, selecting a structural form of the drive axle: selecting a disconnected drive axle or a non-disconnected drive axle according to the type of a vehicle suspension; and S3, after selection of the structural form of the steering drive axle component is completed, design of all the components of the steering drive axle is conducted, and the design comprises design of a main speed reducer, design of a differential mechanism and an axle housing and design of a steering drive axle lifting system. According to the method, firstly, a main speed reducer and a differential mechanism of the amphibious vehicle are selected and calculated, the main speed reducer and the differential mechanism which are stable in power and stable in transmission are designed under the conditions that the specified size is small and the vehicle space is saved, and secondly, in the aspect of a vehicle suspension, a double-wishbone suspension and a damper with a lifter and an air spring are selected through analog calculation; the manufacturing cost is fully saved, and the design requirement is met.

Description

Technical Field

[0001] The invention belongs to the technical field of amphibious vehicle research, and in particular relates to a design method for a liftable steering drive axle of an amphibious vehicle. Background Art

[0002] Amphibious vehicles are special vehicles that have the ability to travel on land and water, and have the performance of both vehicles and ships, and can enter water and land. Because they can operate in various complex terrain environments, they are often used in emergency rescue fields. [2] . In recent years, my country's economy has developed by leaps and bounds, and the development of the automotive field is booming. With the rapid development of the automobile manufacturing industry, the research and development of amphibious vehicles is also booming. However, it is still in a passive and backward state of "pale and weak research and development", and there are very few companies in my country with mass production capabilities. In order to meet the requirements of emergency management departments in rescue and disaster relief, it is now necessary to design the steering drive axle of amphibious vehicles. Through this design, the designed steering drive axle is simple, safe, and more cost-effective.

[0003] At present, the main amphibious vehicles in my country, except for military vehicles, are almost all imported from foreign countries, including light amphibious vehicles used for rescue and disaster relief. To improve the market competitiveness of Chinese automobiles, the steering drive axle is very important in the whole vehicle.

[0004] Therefore, we propose a design method for a liftable steering drive axle of an amphibious vehicle to solve the above problems. Summary of the invention

[0005] The technical problem to be solved by the present invention is to provide a design method for a liftable steering drive axle of an amphibious vehicle.

[0006] The objective of the present invention is achieved through the following technical solutions:

[0007] A design method for a liftable steering drive axle of an amphibious vehicle comprises the following steps:

[0008] S1. Selection of drive axle structure: According to the type of vehicle suspension, choose to use a disconnecting drive axle or a non-disconnecting drive axle structure design;

[0009] S2. Selection of the structural form of the steering drive axle components:

[0010] S2.1、Selection of the structural form of the main reducer;

[0011] S2.2, selection of differential structure;

[0012] S2.3, drive wheel transmission mode selection;

[0013] S2.4, selection of wheel lifting system;

[0014] S3. After the structural form of the steering drive axle components is selected, the design of each component of the steering drive axle is carried out, including the design of the main reducer, the design of the differential and the axle housing, and the design of the steering drive axle lifting system.

[0015] Furthermore, the main reducer design includes:

[0016] (1) Selection of the reduction form of the main reducer and the number of teeth Z1 and Z2 of the main and driven gears;

[0017] (2) Determination of gear calculation load;

[0018] (3) The pitch circle diameter D2 and end face module m of the driven bevel gear of the main reducer s Selection and verification;

[0019] (4) Selection of tooth width b1 and b2 of the driving and driven bevel gears;

[0020] (5) Selection of midpoint helix angle β and normal pressure angle α;

[0021] (6) Strength calculation of main reducer bevel gear;

[0022] (7) Calculation of load on the bevel gear bearing of the main reducer;

[0023] (8) Bevel gear material selection: carburized alloy is selected.

[0024] Furthermore, the design of the differential and the axle housing includes the following steps:

[0025] (1) Design of bevel gear differential;

[0026] (2) Calculation of differential gear strength;

[0027] (3) Design of wheel transmission device;

[0028] (4) Spline stress verification;

[0029] (5) Drive axle housing design;

[0030] (6) Bridge shell stress analysis and strength calculation.

[0031] Furthermore, the design of the driving axle lifting system includes the following steps:

[0032] (1) Design of vehicle suspension;

[0033] (2) Design of elastic elements;

[0034] (3) Selection and design of shock absorbers;

[0035] (4) Design of lifting system;

[0036] (5) Feasibility test of lifting system.

[0037] Furthermore, the reduction form of the reducer and the selection of the number of teeth Z1 and Z2 of the master and slave gears: its transmission mode uses spiral bevel gear transmission. When determining the master and slave gears, the number of teeth Z1 and Z2 cannot have a common divisor; the sum of the number of teeth is ≤40; Z1 is usually ≥9; according to the above requirements, take:

[0038] The number of teeth on the driving wheel is Z1=13, the number of teeth on the driven wheel is Z2=48, the main reduction ratio is i0=3.7, and the gear ratio is

[0039] Furthermore, the feasibility test of the lifting system includes the following:

[0040] (1) Calculation of wheel lifting height;

[0041] (2) Calculation of telescopic travel of the lifter;

[0042] (3) Calculation of double wishbone length;

[0043] (4) Installation design of lifting system.

[0044] Compared with the prior art, the present invention has the following beneficial effects:

[0045] The present invention firstly calculates the selection of the main reducer and differential of the amphibious vehicle, and designs a main reducer and differential with stable power and smooth transmission by keeping the specified size small and saving vehicle space. Secondly, in terms of vehicle suspension, through simulation calculation, a double wishbone suspension and a shock absorber with a lifter and an air spring are selected, which fully saves manufacturing costs and meets design requirements. BRIEF DESCRIPTION OF THE DRAWINGS

[0046] Figure 1 It is a schematic diagram of the disconnect drive axle structure;

[0047] Figure 2 It is the force diagram of the tooth surface of the active bevel gear;

[0048] Figure 3 This is a schematic diagram of cantilever support;

[0049] Figure 4 This is the bearing arrangement dimension drawing of single-stage main reducer;

[0050] Figure 5 This is a simplified diagram of the lifting system;

[0051] Figure 6A diagram showing the wheel rising;

[0052] Figure 7 This is a simplified diagram of the wheel lowering. DETAILED DESCRIPTION

[0053] The following will describe the implementation methods of the present application in detail with the help of accompanying drawings and examples, so that the implementation process of how the present application uses technical means to solve technical problems and achieve technical effects can be fully understood and implemented accordingly.

[0054] The invention provides a design method for a liftable steering drive axle of an amphibious vehicle. The design is carried out according to the following design requirements of the amphibious vehicle: vehicle-vehicle integrated structure, jet pump propulsion, comprehensive capabilities of multiple working conditions such as rapid water and land rescue, towing rescue and water pump fire fighting, curb weight ≤1.95t, carrying ≥9 persons, effective load ≥1t, rescue towing force ≥1.5t, maximum wave height 2.5m, maximum wind resistance level 7, automatic righting function after capsizing 180°, automatic righting time in flipped state not exceeding 60s, water speed ≥55km / h, hull vertical impact acceleration resistance ≥2.5g, land speed ≥85km / h, approach angle and departure angle ≥30 degrees.

[0055] Selection of drive axle structure

[0056] The drive axle is classified into two types: disconnecting drive axle and non-disconnecting drive axle. The choice of drive axle is mainly based on the type of vehicle suspension. If the vehicle is designed with independent suspension, a disconnecting drive axle is used. Otherwise, a non-disconnecting drive axle is selected.

[0057] The disconnected drive axle used in this design is characterized by the drive axle housing and beam being disconnected, rather than being connected to the drive wheels as a whole. Figure 1 As shown, the wheels are connected to the vehicle body through independent suspension, which can prevent the vehicle transmission and steering drive axle guide device from interfering with each other when the wheels on both sides are subjected to uneven forces and on uneven roads.

[0058] Structural selection of steering drive axle components

[0059] The main components of the main reducer in the steering drive axle of an amphibious vehicle. In order to adapt to the driving and mission requirements on land and in water, when driving on various roads, the main reducer is required to have the required driving force and speed. At the same time, after selecting the design standard of the main reducer, the differential, wheel transmission, guide mechanism, and wheel lifting must meet the mission requirements. Within a reasonable range, the design should be reasonable in structure and meet the requirements of vehicle driving in water.

[0060] Main reducer structure

[0061] When choosing the structural form of the main reducer of an amphibious vehicle, the type of gear and the form of reduction must first be considered.

[0062] Under the specified dimensions of the amphibious vehicle, the gear transmission mode of the main reducer is spiral bevel gear transmission. Figure 2 As shown in the figure, its main structural feature is that the axes of the driving bevel gear and the driven gear intersect vertically at one point.

[0063] In order to meet the use requirements of amphibious vehicles, the structural selection of the main reducer is particularly important. This design adopts a single-stage main reducer. It has the advantages of simple structure, small mass and compact structure.

[0064] Differential structure

[0065] The differential of an amphibious vehicle mainly distributes torque between two output shafts and ensures that the output shafts at both ends can rotate at different angular velocities. After verification of the scheme, the amphibious vehicle adopts symmetry. It has the advantages of simple structure and light weight, so it is selected for this design.

[0066] Drive wheel transmission mode selection

[0067] After the above-mentioned disconnected drive axle configuration is selected, a reasonable transmission method must be selected to transfer the torque from the differential half-axle to the drive wheels. Since the wheels are raised and lowered in this design, the suspension and transmission synchronous lifting and lowering issues must be considered. Secondly, this design refers to the off-road vehicle standard and adopts an independent suspension, so the drive wheel transmission method is selected as a universal joint transmission device.

[0068] Wheel lift system options

[0069] This design is mainly to meet the rescue mission of amphibious vehicles. In order to reduce the driving resistance of the vehicle in the water, this design installs a wheel lifting system on the steering drive axle of the vehicle. The air spring is an elastic element with good vibration isolation. It is the core component of the air suspension system. It is composed of a cover plate, a bladder, a base, a buffer block and other parts. Compared with ordinary suspensions in the past, air suspension can extend the service life of the vehicle, buffer the impact force brought by the road surface, and attenuate the vibration caused by it, improve the comfort of the whole vehicle, and greatly reduce the damage of the vehicle to the road surface. In the field of commercial vehicles, replacing leaf springs with air springs is an inevitable trend of development and an objective requirement of reality.

[13] After theoretical demonstration and repeated experiments, we chose to use air springs and hydraulic compression cylinders to lift the vehicle.

[0070] Main reducer design

[0071] The design calculation of the main reducer of an amphibious vehicle mainly includes the selection of the acceleration form of the main reducer, the determination of the gear load and the calculation of strength. The focus is to design a main reducer that meets the requirements of this time by considering the special properties of the vehicle under the traditional design method.

[0072] The reduction form of the main reducer and the selection of the number of teeth of the main and driven gears Z1 and Z2

[0073] The amphibious vehicle design uses a single-stage main reducer, which has the advantages of low manufacturing cost and high transmission efficiency. The transmission method uses spiral bevel gear transmission.

[0074] Secondly, when determining the driving and driven gears, the number of teeth Z1 and Z2 cannot have a common divisor; the sum of the number of teeth is ≤40; at the same time, in order to ensure that the gears have high fatigue strength, low working noise and relatively stable gear meshing. Z1 is usually ≥9; according to the above requirements, take:

[0075] The number of teeth on the driving wheel is Z1=13, the number of teeth on the driven wheel is Z2=48, the main reduction ratio is i0=3.7, and the gear ratio is Determination of calculated load on gears

[0076] First, calculate the torque T ce , from the formula And solve the given data for this design.

[0077] T ce ——Represents the calculated torque (N·m);

[0078] T emax ——The maximum torque of the engine; T emax =300N·m;

[0079] n——is the number of drive axles, from the drive scheme FR we know that: n=1;

[0080] i1——represents the first gear ratio of the transmission; i1=3.45;

[0081] i f ——is the transfer case transmission ratio; from the drive scheme FR, we know that: i f =1;

[0082] i0——represents the transmission ratio of the main reducer; i0=3.7;

[0083] η——Transmission efficiency; η=0.96;

[0084] k——represents the torque conversion coefficient of the torque converter, k=[(K0-1 / 2)]+1, where K0 represents the maximum torque conversion coefficient; when selecting a normal transmission system: k=1;

[0085] K d——dynamic load coefficient; select K d =1.

[0086]

[0087] Next, calculate the torque T using the above formula: cs :

[0088]

[0089] Since this design is an amphibious vehicle, it is assumed that there is no drive wheel slippage. Even if there is slippage, the output torque will still be smaller than the calculated torque of the driven bevel gear.

[0090] The calculated torque of the active bevel gear designed this time is:

[0091]

[0092] Where: η G ——Indicates transmission efficiency, η G =95%. i0——main reduction ratio.

[0093] When calculating the maximum stress in this design, the torque selection should be the minimum value of the previous two calculation results, that is, T c = min[T ce ,T cs ], from the above data we get T c =6114.46N·m,

[0094] The pitch circle diameter D2 and end face module m of the driven bevel gear of the main reducer s

[0095] Now select and check the driven bevel gear D2 and end face module.

[0096] D2 is initially selected according to the formula, namely:

[0097]

[0098] In formula (3-4), the diameter coefficient is Its value range is: 13.0~15.3, and 15 is selected in this design; T c is the calculated torque (N·m), T c = min[T ce ,T cs ].

[0099] Gear end face module m of driven bevel gear s Calculated by the following formula:

[0100]

[0101] At the same time s Should meet:

[0102]

[0103] Round to get m s =6.

[0104] In formula (3-6), K m The value range of is 0.3~0.4, and 0.35 is selected in this design. The diameter D1 of the large end pitch circle of the active bevel gear is calculated as follows:

[0105] D1=m s × Z1=6×13=78 (3-7)

[0106] The tooth widths b1 and b2 of the tooth surface of the driving and driven bevel gears

[0107] The bevel gear tooth surface designed in this project cannot be too wide, which will increase the stress concentration of the gear tooth top fillet, which not only reduces the root circle radius but also shortens the service life. The size of the wide surface b2 of the driven bevel gear is:

[0108] b2=0.155D2=0.155×274=42.5≈43 (3-8)

[0109] For the designed arc bevel gear, b1 is set to be 10% larger than b2, so b1=1.1b2=46.7≈47.

[0110] Midpoint helix angle β and normal pressure angle α

[0111] The helix angle of the spiral bevel gear of the main reducer of the amphibious vehicle ranges from 35° to 40°. This design uses a larger β value to prevent excessive axial force, which is 37°.

[0112] At the same time, in order to ensure the main reducer runs smoothly and with less noise, the normal pressure angle is selected as 16° for the spiral bevel gear designed this time.

[0113] Strength calculation of bevel gear of main reducer

[0114] For the strength calculation of the bevel gear of the main reducer of an amphibious vehicle, the circumferential force per unit tooth length must first be calculated. The formula is:

[0115]

[0116] When calculated according to the maximum torque of the engine:

[0117]

[0118] In current actual automobile design, [p] is sometimes 20% to 25% higher than the value in the table.

[0119] Secondly, calculate the bending strength of the gear teeth, and the bending stress at the tooth root is:

[0120]

[0121] Known m s =4mm, so k s =(m s / 25.4) 0.25 =0.63, take k m =1.25, take J w =0.23×0.89=0.205.

[0122] Active bevel gear teeth:

[0123]

[0124] Driven bevel gear teeth:

[0125]

[0126] The tooth surface contact stress of the bevel gear teeth of the main reducer is calculated as:

[0127]

[0128] Among them, C p Pick J J Select J J =0.12.

[0129] Through the above calculation min[T ce , T cs The maximum contact stress calculated is no more than 2800MPa. Therefore, it is shown that the contact stress of the tooth surface of the main and driven gears of the bevel gear of the main reducer of the amphibious vehicle is the same after calculation.

[0130] Load calculation of bevel gear bearings in main reducer

[0131] For the load calculation of the bevel gear bearing of the main reducer, the circumferential force F at the center of the tooth width must first be calculated as:

[0132]

[0133] Determined by the above formula, that is:

[0134] D m2 =D2-b2sinγ2 (3-16)

[0135] From the calculation, we can know that the design selected is a spiral bevel gear pair, so the circumferential force on the driving and driven gears is equal.

[0136] like Figure 2 The figure shows the force diagram of the active bevel gear tooth surface. It can be seen from the figure that its spiral direction is left-handed, and its rotation direction is counterclockwise when viewed from the top of the cone.

[0137] Depend on Figure 2 It can be seen that:

[0138] F=F T cosαcosβ (3-17)

[0139] F N =F T sinα=Ftanα / cosβ (3-18)

[0140] F s =F T cosαsinβ=Ftanβ (3-19)

[0141] From the calculation, we know that the axial force F az and radial force F Rz They are

[0142] F az =F N sinγ+F s cosγ (3-20)

[0143] F Rz =F N cosγ-F s sinγ (3-21)

[0144] For the shaft design, since the main reducer input shaft is a gear shaft such as Figure 3 As shown, the material of the shaft is the same as that of the active drive, 20CrMnTi, carburized and quenched, σ b =1100MPa;σ s =850MPa;σ -1 =490MPa; τ -1 =526.5MPa; [σ -1 ]=70MPa; take A=105, shaft power: P=P w η=71×0.96=68.16kW, minimum diameter of the shaft:

[0145]

[0146] The support form of the active bevel gear in this design is cantilever support. In this design, it is considered to reduce the length of the cantilever a and increase the distance b, so as to improve the strength of the support.

[0147] At the same time, the design must consider increasing its support stiffness, so the support distance b>2.5a, and at the same time greater than 70% of the gear pitch diameter.

[0148] From the previous calculation, we know that the large end pitch circle of the main reducer driving bevel gear is 52mm, so the support distance is:

[0149] b≥0.7×52=36.4mm (3-23)

[0150] Bevel gear length l c It is approximately = b1×cosγ1=30.301×cos15.15°=29.248mm.

[0151] The cantilever length a should be taken from the bearing in the middle of the gear, so a=14.624mm.

[0152] In order to facilitate the disassembly and assembly of the bearing near the gear, a shoulder should be made between the gear and the bearing, i.e. V-VI =11mm.

[0153] The cantilever length a should be a=14.624+7=21.624mm, taking a=24mm. Taking b=70mm, then 2.5a=2.5×22=55mm≤b=70mm, which is reasonable.

[0154] The bearing far from the gear is initially selected as a tapered roller bearing of model 30209, with an inner diameter and width of d = 45 mm and B = 19 mm. II-III =45mm, I II-III =19mm.

[0155] In order to make this design easy to install, a larger bearing should be selected near the gear. A 30210 tapered roller bearing with an inner diameter and width of d = 50 mm and B = 20 mm should be selected. IV-V =50mm, I IV-V =31mm; According to the installation dimensions of 30210 bearing, take d V-VI =55mm.

[0156] Because the big end of the 30210 bearing faces outward, the inner ring is directly positioned by the coupling. Therefore, in order to ensure that the bearing is not over-positioned, take: d III-IV =d II-III =45mm. Calculated

[0157] I III-IV =bIII-III -I IV-V =70-19-31=20m (3-24)

[0158] The initial spline module is m=1.5, the number of teeth is z=28, the working length is 50mm, and the length of the threaded connection shaft is 30mm.

[0159] Take the middle section of the gear to the center of the large shaft as

[0160] Take the distance between the centers of the two bearings

[0161] Take the distance from the center of the small bearing to the end as

[0162] Perform force analysis on the input shaft, M z It is divided into horizontal and vertical bending moments. z =-64526N.mm, M on vertical surface z =0N.mm. Calculate the horizontal support reaction force, assuming F H2 Left, F H1 Go right, by F H2 R2=F H1 R1+M a have to:

[0163]

[0164] F H1 =F H2 -F az , where R1 = 22.5 mm, R2 = 25 mm, then F H1 =4201.7N, F H2 =-1200.5N.

[0165] Vertical support reaction, assuming F V2 Up, F V1 Downward, by F V2 L1-F V1 (L1+L2), F V1 =F V2 -F Rz Get F V1 =-329N, F V2 =3h27N.

[0166] Horizontal bending moment, M H2 =F H2 ×R2=-30012N,M H1 =F H1×R1=94539N,M z =-64526N.

[0167] Vertical bending moment: M V1 =F V1 ×(L1+L2)=157403N,M V2 =-157403N

[0168] Driving gear torque: T = 474.13N.mm.

[0169] Resultant bending moment:

[0170] The torque of this design is calculated according to the pulsating cycle change. Taking a = 0.6, the bending moment at point 1 is:

[0171]

[0172] The bending moment at 2 points is:

[0173]

[0174] To check the data,

[0175]

[0176] The calculation results show that the designed shaft meets the strength requirements.

[0177] Depend on Figure 4 The figure shows the layout of the dimensions of the cantilever support, and the selection of the load calculation formula for each bearing is shown in Table 3-2.

[0178] After the load on the bearing is calculated and determined, the bearing model is selected by calculating its life.

[0179] The calculation formula for the load on the bearing is shown in Table 3-1:

[0180] Table 3-1 Loads on bearings

[0181]

[0182] Known parameter L h =20×20×24=14400h, n=4500r / min.

[0183] From the above data, calculate the torque T on the main reducer driving bevel gear dz for:

[0184]

[0185] Where: T emax—— is the maximum torque; T emax = 500 N·m

[0186] f iI , f iII , f iIII , f iIV —— The utilization rate of the transmission in each gear; f iI = 1, f iII = 4, f iIII = 20, f iIV = 75

[0187] i gI , i gII , i gIII , i gIV —— The transmission ratios of the I, II, III, IV gears and reverse gear of the transmission; i gI = 3.45, i gII = 2.62, i gIII = 1.71, i gIV = 1

[0188] f TI , f TII , f TIII , f TIV —— The engine torque utilization rate when the transmission is in each gear, f TI = 70, f tII = 65, f TIII = 60, f TIV = 60

[0189] Therefore, through calculation, it can be known that:

[0190]

[0191] For the preliminary selection of bearing B, the tapered roller bearing 30210 is selected, C = 73.2 KN, F a / F r = 0.31 < e = 0.4, so X = 1, the load factor is 1.2, and the equivalent load P = f d F r = 3854 N. For ball bearings, Then the basic rated life of the bearing The service life of the designed vehicle in this design is 20 years. Calculated by 20 days per year and 24 hours of work per day, the working hours are: 30 × 20 × 24 = 14400 h. Therefore, L h = 94287.6 h > 14400 h, meeting the requirements.

[0192] The material of bevel gears

[0193] The working conditions of the bevel gears of the drive axle designed this time are relatively harsh compared to the traditional ones. The working conditions of the bevel gears designed this time are characterized by high load pressure, long working hours, relatively complex working conditions, and impact resistance, which are the weak links of the transmission system in the design of this amphibious vehicle. Therefore, the material of the bevel gears must meet the working requirements. The bevel gears of the main reducer are made of carburized alloy, and this design uses 20CrMnTi with nitriding and quenching.

[0194] Design of differential and axle housing

[0195] The differential design of amphibious vehicles is particularly important for vehicles. The main function of the car differential is to transmit power to the half-axles on both sides of the differential, so that the half-axles on both sides rotate at different speeds, so that the wheels on both sides of the car can move with unequal spacing as much as possible, reducing the friction of the tires on the ground.

[0196] Differential type selection

[0197] The traditional differentials are mainly in the form of symmetrical bevel gear differential and ordinary bevel gear differential. The widely used differential in automobiles is the symmetrical bevel gear differential, which has the characteristics of reasonable structure and small mass. Therefore, the ordinary bevel gear differential is selected in this design.

[0198] According to the motion analysis, it can be calculated:

[0199] ω1+ω2=2ω0 (4-1)

[0200] According to the moment balance, it can be calculated:

[0201] T1+T2=T0 (4-2)

[0202] T2-T1=T r (4-3)

[0203] In this design, the locking coefficient k is used to represent the differential performance, which can be determined as:

[0204] k=T r / T0 (4-4)

[0205] Combining the above formula with the moment balance formula, we can get:

[0206]

[0207] From the calculation, it is known that the torque ratio of the half shaft is k b =T2 / T1, so k b Calculated between k:

[0208]

[0209] Through scheme comparison, this design selects k as 0.1, that is, kb is 1.22. At the same time, the counter torque of the left half shaft to the differential is T1 = 748.35N·m, and the counter torque of the right half shaft to the differential is T2 = 914.65N·m. The friction torque inside the differential is T r =166.3N·m.

[0210] Design of differential for ordinary bevel gears

[0211] Since the design is an amphibious steering drive axle differential, the number of planetary gears n is set to n=2 according to the design requirements.

[0212] According to the design formula:

[0213]

[0214] K b ——Represents the spherical radius coefficient of the planetary gear, K b =2.5~3.0. Since the amphibious vehicle designed this time has two planetary gears, the maximum value is taken, K b =3.

[0215] T d ——represents the calculated torque of the differential (N·m), T d = min[T ce ,T cs ]=1663.13N·m. After this calculation, we get: R b =35.54mm. The pitch distance A0 = (0.98-0.99)R b , select the coefficient as 0.99, then A0=35.19mm.

[0216] Secondly, for the number of teeth z1 and z2 of the planetary gear and the half-axle gear designed in this design, according to the working requirements, the number of teeth z1 of the planetary gear is 11, and the number of teeth z2 of the half-axle gear is 22. z2 / z1=2, within the range of 1.5 to 2, which meets the design requirements. According to the above size selection, the pitch cone angle γ1, γ2 and module m of the planetary gear and the half-axle gear are:

[0217] γ1=arctan(z1 / z2)=15°15' (4-8)

[0218] γ2=arctan(z2 / z1)=74°85' (4-9)

[0219] The end face module of the large end of the bevel gear can be calculated by the formula: m = 2.86, rounded to 3. Then: the large end pitch circle diameter of the planetary gear d1 = z1m = 33mm, then the large end pitch circle diameter d2 = z2 = 66mm.

[0220] This design selects a differential gear with a pressure angle of α=22.5deg and a tooth profile with a tooth height coefficient of 0.8.

[0221] From the above data, the diameter d of the planetary gear shaft can be obtained as:

[0222]

[0223] Where: T0——The torque transmitted by the differential case designed this time, T0=T d =min[Tce,Tcs]=1663.13N·m,[σ c ] is the allowable extrusion stress of the bearing surface designed in this design, which is 98MPa. n——is the number of planetary gears in this design, n=2. d ——represents the distance from the midpoint of the planetary gear support surface to the cone top, r d =0.5=d2'=0.5×0.8×d2=26.4mm.

[0224] From the above calculation, we know that d = 17.09 mm, so we take d = 17 mm. Then the support length L = 1.1d = 18.7 mm.

[0225] Strength calculation of differential gears

[0226] When an amphibious vehicle is turning, the gear pressure is high, so the bending strength must be calculated. Then the gear bending stress σ w for:

[0227]

[0228] T c ——Indicates the calculated torque of the half-axle gear. When T0=min[T ce , T cs ], T c =0.6×T0=997.88Nm

[0229] K s ——K s =1.

[0230] K m ——Indicates the tooth surface load distribution coefficient. K m =1~1.25, choose K m =1.25.

[0231] K v ——Represents the quality coefficient, K v =1.

[0232] m——is the end face modulus of this design, m=3.

[0233] b2——is the tooth width of the half-shaft gear. The tooth width of the bevel gear designed this time is 0.3 times of its pitch cone moment A0, that is, b2=0.3A0=10.56mm.

[0234] d2——represents the pitch circle diameter of the large end of the half-shaft gear, d2=66mm.

[0235] J——is the comprehensive coefficient, J=0.2265.

[0236] From the above data, we can calculate that when T0=min[T ce , T cs ], [σ w ]=980MPa,σ w =263.47MPa<[σ w ], which meets the design requirements.

[0237] Wheel transmission design

[0238] According to the load-bearing and actual situation of this design, the design selects a full-floating half-axle. Its main structural feature is that the flange at the outer end of the half-axle is connected to the wheel hub by screws, and the wheel hub is supported on the half-axle sleeve of the drive axle housing based on two tapered roller bearings.

[0239] The diameter of the semi-axle rod in this design is initially selected as:

[0240]

[0241] Among them, d is the diameter of the full-floating half-shaft rod designed in this design; Represents the calculated torque of the half shaft; K is the diameter coefficient, which is usually in the range of 0.205 to 0.218, and is taken as 0.21 in this design.

[0242] The calculated load of the full floating half axle can be calculated according to the wheel adhesion moment The calculation can be obtained:

[0243]

[0244] Among them, T emax is the maximum torque of the engine, i0 is the main reducer transmission ratio, i1 is the first gear ratio of the transmission, i f is the differential transmission ratio, ξ is the torque distribution coefficient of the differential, which can be 0.6 for the conical planetary gear differential. Substituting the data into the data, we get:

[0245]

[0246] Then d = 21.27 mm,

[0247] Ip is the polar moment of inertia of the semi-axial section, I p =πd 3 / 32. After verification, the torsional shear stress of the half shaft should be 500-700MPa, and τ=509.63MPa meets the requirements.

[0248] Spline stress calculation

[0249] For the stress calculation of the input shaft spline of this design, the coefficient ψ is introduced to calculate the influence of the uneven distribution of the actual load on each spline tooth. At the same time, the strength condition of the spline connection is: the spline is 30 degrees, the module m=1.5, and the number of teeth z=28.

[0250] When it is a static connection:

[0251]

[0252] When it is a dynamic connection:

[0253]

[0254] Where: ψ is the coefficient of uneven load distribution, and ψ is taken as 0.8.

[0255] z – the number of teeth on the spline.

[0256] l——working length of tooth, mm;

[0257] h——working height of the spline tooth side. For involute splines, α=30°, h=m, where m is the module;

[0258] d m ——involute spline, d m =mz=1.5×28=42mm.

[0259] The half shaft and half shaft gear of the amphibious vehicle are designed to be connected by involute spline. At the same time, the spline should be checked for extrusion stress and key tooth shear stress. The extrusion stress is not more than 200MPa, and the shear stress is ≤73MPa. The spline degree is 30 degrees, the module m=1.5, and the number of teeth z=15.

[0260] The shear stress τ of the half shaft spline designed in this paper s for:

[0261]

[0262] Then the half shaft spline extrusion stress σ c for:

[0263]

[0264] Where: T is the maximum torque of the half-axis designed in this design, T = 1039.46Nm;

[0265] D B ——Represents the outer diameter of the half shaft spline, D B =m(Z+1)=1.5×(15+1)=24mm;

[0266] d A ——Represents the inner diameter of the matching spline hole:

[0267] d A =m(Z+1.5)=1.5×(15+1.5)=24.75mm (4-17)

[0268] Z——The number of spline teeth;

[0269] L P ——The working length of the spline designed this time, L P =105mm;

[0270] b——spline tooth width, b=πm=π×1.5=4.71mm;

[0271] φ——is the uneven coefficient of load distribution, which is selected as 0.75 in this calculation.

[0272] Drive axle housing design

[0273] Drive axles can be classified in many forms. For example, according to the form of the axle housing, there are disconnected and integral types. Disconnected types are mostly configured in passenger cars with independent suspension, and integral types are mostly used in commercial vehicles. The choice here is mainly based on the balance between cost and function. Under the same load-bearing capacity, the cost of using independent suspension is much higher than that of non-independent suspension.

[15] Since this design is an amphibious vehicle, there are certain requirements for the rigidity of the bridge housing, the disassembly, adjustment, repair and maintenance of the main reducer and differential.

[0274] Bridge shell stress analysis and strength calculation

[0275] When the traction or braking force of the amphibious vehicle reaches the maximum, the torsional stress τ and bending stress σ of the dangerous end face of the bridge shell are:

[0276]

[0277]

[0278] Where: M V ——M v =F x2 ×b;

[0279] b——indicates the distance from the wheel surface to the bridge housing leaf spring seat in mm;

[0280] M h ——W h =F x2 ×b;

[0281] T t ——Torque on the dangerous section, T t =F x2 × r ;

[0282] W V , W h ——represents the bending section coefficients of the vertical plane and horizontal plane bending of the dangerous section of the bridge shell respectively;

[0283] W T ——It represents the torsional section coefficient of the dangerous section.

[0284] When the lateral force of the vehicle is the largest, the bending stress σ of the section at the inner and outer leaf spring seats of the bridge housing is i ,σ0 should be:

[0285]

[0286]

[0287] Among them, F z2i 、F z20 is the vertical reaction force of the inner and outer wheels on the ground; r r is the rolling radius of the wheels of the amphibious vehicle; is the sideslip coefficient when the vehicle slides sideways.

[0288] When a vehicle passes through an uneven road surface, the bending stress of the dangerous section for:

[0289]

[0290] It is known that the permissible bending stress of the bridge housing of the steering drive axle designed this time is 300-500Mpa, and the permissible torsional shear stress is 150-400Mpa. Therefore, the forgeable cast iron bridge housing is designed to have a smaller value, and the steel plate stamping bridge housing is designed to have a larger value.

[0291] Design of the Lifting System of the Steering Drive Axle

[0292] Vehicle suspension design

[0293] The quality of suspension performance is the main factor affecting the mobility of off-road vehicles. With the development of the times, modern off-road vehicles have put forward higher requirements for suspension performance. The traditional non-independent suspension has seriously restricted the improvement of off-road vehicle's smoothness, handling stability and passability.

[16] .

[0294] In order to solve the above problems and meet the needs of amphibious vehicle rescue missions, a double wishbone independent suspension suitable for amphibious vehicles is now designed.

[17] The simulation and test results were used to verify that the nonlinear stiffness mathematical model can accurately calculate the dynamic characteristics of the suspension stiffness, providing a theoretical basis and reference for the accurate design of the suspension stiffness. At the same time, it was determined that the double wishbone independent suspension is suitable for the driving and mission requirements of amphibious vehicles. The length of the two upper and lower swing arms of the double wishbone independent suspension can be equal or unequal. According to the design requirements of this time, unequal length swing arms are used.

[0295] Design of elastic elements

[0296] In accordance with the design requirements, the elastic element of the suspension is selected as a gas spring. Due to its special structure and gas-solid coupling characteristics, the rubber air spring has better shock absorption and high-frequency shock absorption performance than steel springs. Compared with metal springs, the advantages of air springs are: the speed changes relatively slowly, the changes when subjected to dynamic excitation are relatively small, and it is easier to control

[18] . The working principle of the air spring is to fill compressed air into its flexible and enclosed space, and at the same time use the reaction force of the compressed air in the rubber airbag as an elastic element of the elastic restoring force. Air springs can be divided into bag type and membrane type. The working principle of the bag type air spring is that it consists of a rubber airbag with cords and compressed air enclosed therein. The inner layer of the airbag is made of rubber with good airtightness, and the outer layer is made of oil-resistant rubber. The airbag is generally made into two sections, but there are also single sections or 3 or 4 sections. The more sections there are, the better the elasticity. There are rigid waist rings between the sections to prevent the middle part from radial expansion and prevent friction between the two sections. The airbag seal is composed of the upper and lower cover plates of the airbag.

[0297] The capsule air spring is easy to manufacture, has a simple structure, low manufacturing cost, relatively simple working conditions, and a long service life. The capsule air spring is the most commonly used in automobiles. Compared with the membrane air spring, the capsule air spring has no throttle hole and piston base, so it can be set at a lower working height within a limited installation height, and can also obtain greater elastic deformation, so the capsule air spring is selected in this design.

[0298] Shock absorber selection and design

[0299] In designing the steering drive axle of an amphibious vehicle, the most important component after selecting the suspension is the shock absorber.

[0300] Compared with the traditional shock absorber design, this design must first consider the waterproof setting of the shock absorber when the vehicle is driving in the water. Since most traditional shock absorbers use oil to lift the vehicle, they must be waterproofed in the water, which greatly increases the design cost. In order to save costs, increase the service life of the shock absorber, and prevent oil from polluting the water, this design will combine the design requirements and actual needs to transform and upgrade the shock absorber.

[0301] According to the requirements of this design, a combination of lifting system and air spring is selected to achieve vehicle shock absorption. When the vehicle is driving on the road, the oil pressure monitoring system is used to open the oil inlet and outlet of the lifting system and adjust the oil pressure of the lifting system according to the vibration of the vehicle, thereby playing the role of vehicle shock absorption.

[0302] Design of lifting system

[0303] In order to meet the requirements of wheel lifting, a hydraulic lifter is designed this time. The working principle is as follows Figure 5 As shown in the figure, the medium used in the hydraulic cylinder is incompressible oil, and its outstanding feature is its high sensitivity. When the hydraulic cylinder is working, the pressure source of the piston in the hydraulic cylinder is: receiving the oil pressure from the upper and lower oil inlets and outlets. According to the vehicle condition, the oil pressure on one side decreases, and the oil pressure on the other side increases, so that the piston in the lifter produces reciprocating telescopic motion, thereby driving the wheels to rise and fall to meet the needs of amphibious vehicle missions.

[0304] The design of this lifting system focuses on steering drive, and its working principle is as follows:

[0305] (I) Working principle of wheel lifting

[0306] By installing a throttle valve in the steering drive axle, the filling and draining speed of the hydraulic cylinder can be controlled, thereby controlling the speed of wheel lifting and lowering. Figure 6 As shown, when the wheel is lowered, a two-position two-way solenoid ball valve is set to control the wheel to be lowered, and the hydraulic oil is controlled to enter from the upper oil inlet to drive the hydraulic cylinder to contract and raise the wheel. At the same time, a safety relief valve is set to limit the upper pressure limit of the oil inlet to protect the entire hydraulic system.

[0307] (II) Working principle of wheel lowering

[0308] When the driver turns on the wheel lowering switch, the two-position three-way solenoid valve is selected as the wheel raising valve. Figure 7As shown, the lower end of the hydraulic cylinder is the oil inlet and the upper end is the oil outlet, which lifts the wheel; at the same time, a one-way valve is installed on the frame to ensure that the hydraulic oil used for lifting can only flow in one direction. When the driver manually drains the oil, it can ensure that the upper and lower oil inlet and outlet circuits share a set of drain pipelines, and the upper and lower oil circuits can be isolated to ensure independent control of the vehicle's lifting and lowering.

[0309] Feasibility test of lifting system

[0310] Wheel lift height calculation

[0311] It is known that the wheel model is 275 / 45R21, that is, the wheel section width is 275mm, the tire section is: H = 45÷100×275=123.75, the wheel diameter: D = 123.75×2+21×25.4≈780mm, and it is known that the vehicle clearance is 250mm, so in order to meet the wheel lifting requirements, the wheel lifting distance is 300mm.

[0312] Calculation of telescopic travel of lifter

[0313] From the above calculation, we know that the wheel lifting distance is 300mm, the lifter rotation position is 1300mm from the ground, the vehicle ground clearance is 250mm, and the kingpin inclination angle is set to 8°. Therefore, the lifter telescopic stroke is calculated as:

[0314] Lifter telescopic stroke L = 300 / cos8°≈303mm (5-1)

[0315] Double wishbone length calculation

[0316] When assembling the steering drive axle, the upper and lower kingpins must be located on the same center line and intersect the center of the universal joint. Through calculation, it is known that the wheel lifting height is 300mm, and the swing arm angle is set to 30°~60°, so the length of the lower swing arm is:

[0317] L1=(300÷2)÷sin30°=300mm (5-2)

[0318] From the above calculations, we can know that the kingpin inclination angle is 8°, the vehicle lifting height is 300mm, and the distance c between the intersection of the kingpin extension line and the wheel and the center line of the wheel is in the range of 40 to 60mm.

[0319] Lifting system installation design

[0320] From the above design calculations, we can know the length of the double wishbone, lifter and telescopic stroke. The vehicle lifting process is simulated by CATIA software. The amphibious vehicle is in normal driving state. When the vehicle is driving in the water, the hydraulic control system of the lifter is controlled. The hydraulic oil enters through the lifter oil inlet, the piston rod moves down, the oil outlet opens, and the wheel is slowly lifted by hydraulic action. After a period of action, the wheel rises to the highest point, completing the wheel lifting.

[0321] When the vehicle completes its underwater mission, the hydraulic control system is manipulated to allow the hydraulic oil in the lifter to enter the hydraulic cylinder through the oil inlet at the lower end. The piston moves upward, and the oil outlet at the upper end discharges the hydraulic oil at the upper end of the piston rod. The wheels slowly descend, and after a period of time, the wheels return to their normal land driving height. Through simulation calculations, CATIA is finally used for modeling to obtain the steering drive axle of the amphibious vehicle.

[0322] In summary, the above are only preferred embodiments of the present invention and are not intended to limit the present invention in any form. Any simple modifications or equivalent changes made to the above embodiments based on the technical essence of the present invention fall within the protection scope of the present invention.

[0323] The above description is only a preferred embodiment of the present invention and does not limit the present invention in any form. Any simple modification, equivalent changes and modifications made to the above embodiments based on the technical essence of the present invention are within the scope of the technical solution of the present invention.

Claims

1. A design method for a liftable steering drive axle of an amphibious vehicle, characterized in that: The steps include: S1. Selection of drive axle structure: According to the type of vehicle suspension, choose to use a disconnecting drive axle or a non-disconnecting drive axle structure design; S2. Selection of the structural form of the steering drive axle components: S2.1、Selection of the structural form of the main reducer; S2.2, selection of differential structure; S2.3, drive wheel transmission mode selection; S2.4, selection of wheel lifting system; S3. After the structural form of the steering drive axle components is selected, the design of each component of the steering drive axle is carried out, including the design of the main reducer, the design of the differential and the axle housing, and the design of the steering drive axle lifting system.

2. The design method of the amphibious vehicle liftable steering drive axle according to claim 1, characterized in that: The main reducer design includes: (1) Selection of the reduction form of the main reducer and the number of teeth Z1 and Z2 of the main and driven gears; (2) Determination of gear calculation load; (3) The pitch circle diameter D2 and end face module m of the driven bevel gear of the main reducer s Selection and verification; (4) Selection of tooth width b1 and b2 of the driving and driven bevel gears; (5) Selection of midpoint helix angle β and normal pressure angle α; (6) Strength calculation of main reducer bevel gear; (7) Calculation of load on the bevel gear bearing of the main reducer; (8) Bevel gear material selection: carburized alloy is selected.

3. The design method of the amphibious vehicle liftable steering drive axle according to claim 1, characterized in that: The design of the differential and the axle housing includes the following steps: (1) Design of bevel gear differential; (2) Calculation of differential gear strength; (3) Design of wheel transmission device; (4) Spline stress verification; (5) Drive axle housing design; (6) Bridge shell stress analysis and strength calculation.

4. The design method of the amphibious vehicle liftable steering drive axle according to claim 1, characterized in that: The design of the driving axle lifting system includes the following steps: (1) Design of vehicle suspension; (2) Design of elastic elements; (3) Selection and design of shock absorbers; (4) Design of lifting system; (5) Feasibility test of lifting system.

5. The design method of the amphibious vehicle liftable steering drive axle according to claim 2, characterized in that: The selection of the deceleration form of the reducer and the number of teeth Z1 and Z2 of the driving and driven gears: its transmission mode uses spiral bevel gear transmission. When determining the driving and driven gears, the number of teeth Z1 and Z2 cannot have a common divisor; the sum of the number of teeth ≤40; Z1 is usually ≥9; according to the above requirements, take: The number of teeth on the driving wheel is Z1=13, the number of teeth on the driven wheel is Z2=48, the main reduction ratio is i0=3.7, and the gear ratio is 6. The design method of the amphibious vehicle liftable steering drive axle according to claim 1, characterized in that: The feasibility test of the lifting system includes the following: (1) Calculation of wheel lifting height; (2) Calculation of telescopic travel of the lifter; (3) Calculation of double wishbone length; (4) Installation design of lifting system.

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