Ship-shore dual-purpose crane facing four-level sea waves

Abstract

The invention provides a four-level sea wave-oriented ship-shore dual-purpose crane, which comprises a rail base, a ship-shore dual-purpose crane, a ship-shore dual-purpose crane and a ship-shore dual-purpose crane, the cart walking base is of a triangular truss structure, stable supporting is formed through an inclined steel frame and a horizontal joist, and a trolley walking track is arranged on the upper portion of the base. The trolley walking base is used for supporting and moving the cantilever crane and realizing transverse and longitudinal movement; the arm frame is used for supporting the grabbing mechanism and carrying out stretching and angle adjustment according to operation requirements; the lifting height of the grabbing mechanism is divided into an upper rail section and a lower rail section, the total lifting height covers the carrying requirements under different water levels, and the dynamic load range is calculated through a lifting dynamic load coefficient formula; the sea wave compensation device realizes effective compensation of four degrees of freedom of heaving, pitching, rolling and yawing by independently controlling the displacement of four hydraulic oil cylinders, provides heaving compensation capability and ship shaking compensation capability, and compensates six degrees of freedom in cooperation with movement of large and small vehicles, so that the stability of the crane is kept.

Description

Technical Field

[0001] The present invention relates to the technical field of ship-shore dual-purpose cranes, and in particular to a ship-shore dual-purpose crane capable of facing level four waves. Background Art

[0002] With the rapid development of the global maritime economy, the demand for ship-to-shore cranes is growing. These cranes must not only efficiently move cargo between ports and coasts but also handle complex offshore replenishment missions. Crane efficiency and safety are particularly important during ship-to-ship and ship-to-shore cargo transfers. However, factors such as wave-induced ship motion and water level fluctuations pose numerous challenges to traditional cranes. For example, wave-induced ship rolling can cause cargo to collide or become suspended during transfers between ships, especially in a Class 4 sea state characterized by distinct wave shapes and widespread whitecaps, increasing the risk of equipment damage and reducing operational efficiency.

[0003] During ship-to-shore handling, especially during the flood and dry seasons, the significant changes in water levels place higher demands on the crane's operating space. During the flood season, the water level is close to the shore, resulting in a reduction in the working space (the ship's deck is about 5 meters above the shore), and the crane needs to operate efficiently in a limited space; during the dry season, the water level drops and the working space increases (the ship's deck is about 10 meters below the shore), and the crane needs to adapt to a larger operating range. During ship-to-ship handling, the height difference between the two ship decks can reach 10 meters, and the crane must be compatible with height changes of ±10 meters. Therefore, developing a ship-to-shore dual-purpose crane that can operate stably under different water level conditions is crucial to ensuring the safety and efficiency of cargo handling.

[0004] To address these complex operational requirements, active heave compensation technology has been integrated into the crane design to enhance its stability and safety in adverse sea conditions. This technology adjusts the position of the crane boom and cargo in real time, keeping them relatively stationary in complex sea conditions. This effectively reduces the risk of cargo collision and overhang caused by vessel motion.

[0005] Patent application document CN107500151A discloses an offshore operation crane, including a slewing support platform, a lifting arm, a lifting hydraulic cylinder, a lifting rope, a lifting rope drum and a lifting hook. The upper part of the lifting sleeve is hinged at the front end of the lifting arm, the middle part of the lifting arm is hinged to one end of the swing hydraulic cylinder, and the other end of the swing hydraulic cylinder is hinged to the middle part of the lifting sleeve. One end of the lifting rod is arranged in the lifting sleeve and is driven by the lifting hydraulic cylinder on the lifting sleeve to extend and retract along the lifting sleeve. The other end of the lifting rod is connected to the fine-tuning bracket, and the two sides of the fine-tuning bracket are respectively connected to two fine-tuning hydraulic cylinders. The lower ends of the two fine-tuning hydraulic cylinders are respectively hinged to the two ends of the rotating hydraulic motor. The rotating hydraulic motor is connected to a clamping bracket, and the two sides of the clamping bracket are respectively connected to two clamping hydraulic cylinders. The two clamps are arranged oppositely and the upper ends are hinged on the clamping bracket. The middle parts of the two clamps are respectively hinged to the lower ends of the two clamping hydraulic cylinders and rotate around the clamping bracket hinge point to relatively clamp the buoy or release the buoy in opposite directions. However, existing active heave compensation systems are mostly used for large-scale deep-sea lifting operations. Their equipment is bulky and expensive, making them difficult to apply to ship-to-shore and ship-to-ship handling scenarios where space is limited and cost is sensitive. Summary of the Invention

[0006] In view of the defects in the prior art, the purpose of the present invention is to provide a ship-shore dual-purpose crane that can withstand level 4 waves.

[0007] The ship-to-shore dual-purpose crane for level 4 waves provided by the present invention comprises:

[0008] The track base is fixed to the foundation with bolts, and its gauge and length are adapted to the water level changes during the flood season and the dry season;

[0009] The trolley travel base adopts a triangular truss structure, which is stably supported by oblique steel frames and horizontal support beams. The upper part of the base is provided with a trolley travel track.

[0010] The trolley travel base is equipped with a 360° rotating structure to support and move the arm and achieve horizontal and vertical movement;

[0011] The boom is used to support the grabbing mechanism and can be extended and adjusted in angle according to operational needs. Its extension range is calculated based on the moment balance formula to adapt to the height difference of ±10 meters between the ship and the shore;

[0012] The lifting height of the grabbing mechanism is divided into upper and lower sections. The total lifting height covers the handling requirements under different water levels, and the dynamic load range is calculated using the lifting dynamic load coefficient formula;

[0013] The wave compensation device independently controls the displacement of four hydraulic cylinders to effectively compensate for four degrees of freedom: heave, pitch, roll, and yaw. It provides heave compensation and ship roll compensation capabilities, and cooperates with the movement of the trolley to compensate for six degrees of freedom, thereby maintaining the stability of the crane.

[0014] The trolley walking base is connected to the track base by bolts, the trolley walking base is installed on the upper track of the trolley walking base, the arm is hinged to the trolley walking base through a rotating structure, the grabbing mechanism is suspended at the end of the arm, and the hydraulic cylinder of the wave compensation device is symmetrically arranged at the connection between the trolley walking base and the arm.

[0015] Preferably, the triangular truss structure of the trolley traveling base meets the following design conditions:

[0016] The axial load of the truss members is optimized based on the material strength formula. Calculate the cross-sectional area, where σ is the allowable stress of the material, F is the axial force, and A is the cross-sectional area of ​​the rod;

[0017] The self-weight of the truss is calculated by the formula of equivalent section moment of inertia Calculation to achieve lightweight design, where I i is the section moment of inertia of a single member, L i is the length of a single rod, L total is the total length of the truss;

[0018] The base stiffness is calculated through the node displacement equation Verification is performed to ensure that the deformation under sea state level 4 does not exceed the safety threshold, where δ is the node displacement, E is the material elastic modulus, I is the section moment of inertia, and L is the length of the rod.

[0019] Preferably, the 360° rotation structure of the trolley walking base is realized by the following method:

[0020] The selection of slewing bearings is based on the calculation of the support reaction force distribution, which meets the following requirements: where R max is the maximum bearing reaction force, M 回转 is the rotational torque, D 轴承 is the bearing diameter;

[0021] The power of the slewing drive motor is given by the formula Determine, where T is the rotational resistance torque, ω is the angular velocity, and η is the transmission efficiency.

[0022] Preferably, the telescopic and angle adjustment of the arm is controlled by:

[0023] The telescopic length of the boom is based on the moment balance formula F 伸缩 ·L 臂架 =G 负载 ·H 高度差 , adapted to the ±10m height difference between ship and shore, where F 伸缩 is the boom extension driving force, L 臂架 is the effective length of the boom, G 负载 is the load weight of the grabbing mechanism, H高度差 is the maximum height difference between ship and shore;

[0024] Boom inclination angle through hydraulic cylinder thrust formula Adjust to ensure the positioning accuracy of the gripping mechanism under dynamic load, where F 液压 is the hydraulic cylinder output thrust, M 负载 L is the moment of load on the hinge point of the boom, 力臂 is the length of the hydraulic cylinder force arm.

[0025] Preferably, the lifting dynamic load coefficient of the gripping mechanism is calculated by the following formula:

[0026]

[0027] in, is the minimum dynamic load coefficient, β2 is the speed influence factor, v q is the lifting speed.

[0028] Preferably, the hydraulic cylinder control method of the wave compensation device includes:

[0029] According to the ship's heave displacement signal, the hydraulic cylinder displacement command is generated through the PID control algorithm, where the error integral term where K p , K i , K d are the proportional, integral and differential coefficients respectively, e(t) is the displacement error signal, and e(τ) is the ship heave displacement signal;

[0030] The heave compensation is calculated by the acceleration integral formula Calculate, the compensation error is less than the set threshold, where α(τ′) is the ship heave acceleration;

[0031] Pitch and roll compensation is achieved through the balance equation ∑F of the hydraulic cylinder thrust and the boom inertia moment 液压 r = I α realization, where F 液压 is the output thrust of the hydraulic cylinder, r is the lever arm, I is the moment of inertia of the arm, and α is the angular acceleration.

[0032] Preferably, the overall stability of the crane is verified by the following steps:

[0033] Basic stability: Based on the geometric relationship diagram of the support reaction force, establish the center of gravity coordinate equation Solve the distribution of support reaction force under the critical state of overturning edge, where G i is the weight of each component, x i is the coordinate of the center of gravity of each component;

[0034] Dynamic stability: Combined with the overturning moment formula M under wind load 风=P·A·h, verifying the critical values ​​of wind speed and wind pressure, where P is wind pressure, A is windward area, and h is the effective height;

[0035] Anti-storm stability: through the moment balance equation M under the maximum wind pressure condition in the non-working state 抗倾 ≥M 风 Calculate the anti-overturning coefficient, where M 抗倾 The anti-overturning moment generated by the crane's own weight.

[0036] Preferably, the steel wire rope of the driving device is selected to meet the following requirements:

[0037] The breaking force of the wire rope is calculated by the formula F 破断 =n·F 工作 ·η 滑轮 Calculate, where n is the safety factor, F 工作 is the maximum working tension of the wire rope, η 滑轮 is the pulley efficiency;

[0038] The pulley diameter is determined by the sheave diameter ratio formula D = h·d, where h is the sheave diameter ratio coefficient and d is the wire rope diameter.

[0039] Preferably, the friction resistance of the operating mechanism is calculated in the following manner:

[0040] Trolley running friction resistance: comprehensive rolling friction resistance Bearing friction resistance and slope resistance F 坡道 =G·sinα to derive the total resistance, where μ 滚动 、μ 轴承 are the rolling friction coefficient and the bearing friction coefficient respectively, α is the angular acceleration, D is the pulley diameter, G is the mass of the running mechanism, and d is the wire rope diameter;

[0041] Frictional resistance of the trolley: Based on the weight distribution formula F 摩擦 =μ·G and dynamic load magnification factor C f Correction, where μ is the comprehensive friction coefficient.

[0042] Preferably, the simulation verification of the wave compensation device includes:

[0043] Generate random signals based on the Longuet-Higgins wave superposition model, Simulate the fourth-order sea state, where ζai is the amplitude of the i-th harmonic, n is the total number of harmonics, ωi is the angular frequency, and ∈i is the random initial phase;

[0044] Calculate the amplitude via the Pierson-Moskowitz spectrum Verify that the cargo displacement after compensation is Δs≤0.035m, where S ζ (ω i ) is the spectral density;

[0045] Energy efficiency optimization is achieved through the hydraulic pump speed adjustment formula Q = k·n and the power recovery rate of the energy recovery device Implementation, where Q is the flow rate, k is the pump displacement coefficient, n is the pump speed, P 回收 、P 输入 For recovered and input power.

[0046] Compared with the prior art, the present invention has the following beneficial effects:

[0047] (1) The crane can cope with different working space restrictions during the flood season and the dry season respectively. When the water level is high during the flood season and the deck is about 5 meters above the shore, the crane can achieve efficient operation in a small space. When the water level drops during the dry season and the deck is about 10 meters below the shore, the crane can still maintain high efficiency in a larger operating space, meeting the different water level requirements of ship-to-shore transportation.

[0048] (2) The crane design takes into account the height difference that may exist in ship-to-ship operations, ensuring that cargo can be transferred smoothly when the height difference reaches ±10 meters. This function not only improves the efficiency of cargo transfer between different types of ships, but also ensures the safety and flexibility of operations.

[0049] (3) The crane is equipped with a six-degree-of-freedom wave compensation system, which uses four hydraulic cylinders to achieve compensation for heave, pitch, roll and bow, and cooperates with the operating mechanisms of the trolley and the carriage to achieve compensation for longitudinal and transverse sway. The simulation results of the system show that the crane's hydraulic wave compensation system can operate stably in sea conditions not exceeding level 4 and in sea areas with an ambient temperature of -10℃ to 50℃, significantly reducing the risk of cargo shaking and collision caused by ship movement. BRIEF DESCRIPTION OF THE DRAWINGS

[0050] 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:

[0051] Figure 1 For automated handling cranes;

[0052] Figure 2 It is the walking base of the cart;

[0053] Figure 3 It is a schematic diagram of the working cycle;

[0054] Figure 4 The geometric relationship diagram of the base support reaction force;

[0055] Figure 5 is the support reaction diagram;

[0056] Figure 6 It is a schematic diagram of the force on the drum;

[0057] Figure 7 for the wave compensation hydraulic system;

[0058] Figure 8 This is the system schematic diagram;

[0059] Figure 9 It is the directional valve control signal;

[0060] Figure 10 is the displacement simulation result;

[0061] Figure 11 This is the simulation principle diagram of the hydraulic cylinder position control;

[0062] Figure 12 This is the simulation result of hydraulic cylinder position control;

[0063] Figure 13 It is a simplified diagram of the execution device;

[0064] Figure 14 This is the schematic diagram of the crane claw offset compensation;

[0065] Figure 15 Simulation results for compensating for crane claw offset;

[0066] Figure 16 To simulate ocean wave signals;

[0067] Figure 17 This is the principle diagram of crane heave compensation simulation;

[0068] Figure 18 The simulation results of cargo displacement relative to the ground. DETAILED DESCRIPTION

[0069] The present invention will be described in detail below with reference to 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, for those skilled in the art, several changes and improvements can be made without departing from the scope of the present invention. These all fall within the scope of protection of the present invention.

[0070] Example

[0071] The present invention provides a ship-shore dual-purpose automated crane and hydraulic system for level 4 waves, comprising:

[0072] like Figure 1The automated handling crane includes: a rail base, a trolley walking base, a small car walking base, a boom, a wave compensation device and a grabbing mechanism.

[0073] 1. Track base

[0074] The crane rails, made of U71Mn, are bolted to the base. With a track gauge of 6,000 mm and a length of 10,000 mm, they are used to secure the crane's overall position and, through the gantry travel mechanism, adjust the crane's position to suit different operating scenarios.

[0075] 2. Trolley Traveling Base

[0076] like Figure 2 The structure is made of Q235 steel, with dimensions of 200mm × 200mm and 6mm thick square steel welded together, and measures 4000mm × 6000mm × 2500mm. The base is designed as a triangular truss structure, consisting of two diagonal steel frames and horizontal joists, providing excellent stability and load-bearing capacity to support the main structure of the crane and ensure its stability and reliability during operation.

[0077] Save materials: Since the truss members mainly bear axial forces, the strength of the material can be fully utilized and the material consumption can be reduced.

[0078] Reduced deadweight: Trusses are lighter in deadweight compared to beams of the same size.

[0079] Increased rigidity: Truss structures can provide greater rigidity.

[0080] The trolley running track is fixed on the upper part of the base, and the track spacing is 4000mm.

[0081] 3. Trolley Traveling Base

[0082] Constructed of sturdy Q235 steel plates, it supports and moves the boom, allowing it to be flexibly adjusted to meet handling requirements at varying heights and distances. The base features a swivel mechanism for the boom, allowing for 360° rotation.

[0083] 4. Boom

[0084] The boom supports the grabbing mechanism and provides the required operating height and reach. It can be extended and adjusted in angle to meet operational needs, ensuring safe and efficient cargo handling in a variety of complex working environments. The boom has an outreach of 2500mm and a height of 3000mm.

[0085] 5. Wave compensation device

[0086] Designed to maintain crane stability in complex sea conditions, the wave compensator independently controls the displacement of four hydraulic cylinders, effectively compensating for heave, pitch, roll, and yaw. It provides 1.5 meters of heave compensation and ±15° of ship roll. Combined with the movement of the trolley, it can compensate for six degrees of freedom, ensuring safe cargo handling in adverse sea conditions.

[0087] Basic parameters of crane

[0088] 1. Lifting capacity: 7t;

[0089] 2. Total lifting height: 15m (lifting height on track: 5m, lifting height below track: 10m);

[0090] 3. Outreach 2500m;

[0091] 4. The track gauge of the trolley is 6000mm and the track gauge of the small car is 4000mm;

[0092] 5. The base distance of the large vehicle is 4000mm, and the base distance of the small vehicle is 4000mm;

[0093] 6. Working speed (lifting and lowering speed 60m / min);

[0094] 7. Rotation speed (6r / min);

[0095] 8. Operating speed 60m / min;

[0096] 9. Maximum wheel load 247kN;

[0097] 10. Working cycle Figure 3 ;

[0098] 11. Impact coefficient For adverse effects, take 1 to 1.1;

[0099] 12. Lifting dynamic load coefficient (dynamic effect),

[0100] Lifting state HC2 slight impact, β2=0.34,

[0101] 13. Operational Impact Coefficient

[0102] 14. The hoisting motor operates in S4 mode, with a duty cycle of 40%.

[0103] Overall stability

[0104] like Figure 4The center formed by the four supporting points of the frame is the origin coordinate, and the direction perpendicular to the track pointing to the sea side is the positive direction of the x-axis.

[0105] The weight of the fixed part (non-rotating part) G1 = 2000kg, the rotating part G2 = 45000kg, and the axial pressure N = G1 + G2 = 47000kg;

[0106] M x =2000*g*1085+45000*g*1490*cosψ;

[0107] M y =45000*g*1490*sinψ;

[0108] The plumb bob force of each leg is:

[0109]

[0110] Substituting the values, we have:

[0111]

[0112] Draw an image like Figure 5 .Depend on Figure 5 It can be seen that the maximum pressure occurs at 33.8 degrees, near the top of A, and the maximum support reaction force R zmax =214083.72N(21823kg), the minimum support reaction force is not zero.

[0113] The maximum number of wheels of the crane, the allowable wheel pressure of the terminal under working condition, Considering subsequent processing and function expansion, the number of single-point large wheels is 2.

[0114] 1. Basic stability

[0115] In the case of no wind and full load lifting, (A) Pointing to the sea side;

[0116] Fixed part P G1 =2000kg, center of gravity coordinates (0,0);

[0117] Rotating part P G2 =45000kg, center of gravity coordinates (1.49, 0);

[0118] The machine weighs 47,000 kg, and the center of gravity is 45,000*1.49 / 47,000=1.43;

[0119] Overturning edge AD: The moment is:

[0120] P G (s / 2-x a )-1.5PQ (R max +tS / 2)

[0121] =47000(3-1.43)-1.5×4500(3.5+1.085-3)

[0122] =63091.25>0

[0123] =63t·m

[0124] (B)

[0125] Fixed part P G1 =2000kg, center of gravity coordinates (0,0);

[0126] Rotating part P G2 =45000kg, center of gravity coordinates (0, 1.49);

[0127] The machine weighs 47,000 kg, and the center of gravity is 45,000*1.49 / 47,000=1.43;

[0128] Overturning edge AB: The moment is:

[0129] P G (B / 2-y G )-1.5P Q (R max -B / 2)

[0130] =47000(2-1.43)-1.5×4500(3.5-2)

[0131] =16665kg·m

[0132] =16.67t·m

[0133] Therefore, basic stability is satisfied.

[0134] 2. Dynamic Stability

[0135] Working condition II, with wind, full load and normal working condition, wind load P W =CKpA=1.4×1×250×5.1=0.1785t, calculated wind pressure p=0.625v s 2 The working wind pressure along the coast is 250N / m 2 , wind speed v sII =20m / s, non-working p at sea III =1800N / m 2 , k h =1 (the wind pressure height coefficient is not considered to be 1 in the working state);

[0136] (1) Effective frontal area 3*1.7=5.1m 2 ;

[0137] (2) The wind force coefficient C is taken from the table as 1.4, and the action height h wII =3.5,α II =6° (take car), h is 15m, P a h α =4.5t×tanα II h α =7.0945t·m;

[0138] (A) The boom is vertical and the trolley faces the sea side.

[0139] Fixed part P G1 =2000kg, center of gravity coordinates (0,0);

[0140] Rotating part P G2 =45000kg, center of gravity coordinates (1.49, 0);

[0141] The machine weighs 47,000 kg, and the center of gravity is 45,000*1.49 / 47,000=1.43;

[0142] Capsizing edge AD, wind direction (land to sea) or capsizing edge AB, wind direction (vertical boom);

[0143] When overturning edge AD, ∑M=(S / 2-x G )-1.3P Q (R max +tS / 2)-P WII h WII -P α h α

[0144] =47000(3-1.43)-1.3×4500(3.5+1.085-3)-0.1785×3.5-7.0945

[0145] =47×1.57-1.3×4.5×1.585-0.1785×3.5-7.0945

[0146] =56.7985t·m>0

[0147] When the overturning edge AB is G )-P' WII h' WII -P α h α

[0148] =47×2-CKpAh wII -7.0945

[0149] =94-1.4×1×250×6.863×10 -3 ×3.5-7.0945

[0150] =78.5t·m>0

[0151] (B) Arm parallel to trolley track

[0152] Fixed part P G1 =2000kg, center of gravity coordinates (0,0);

[0153] Rotating part P G2 =45000kg, center of gravity coordinates (1.49, 0);

[0154] The machine weighs 47000kg, the center of gravity is 0, 45000*1.49 / 47000=1.43;

[0155] Capsizing side AB, wind direction (blowing from back to front along the boom) or capsizing side AD, wind direction (from land to sea perpendicular to the boom); Capsizing side AB:

[0156] ∑M=P G (B / 2-y G )-1.3P Q (R max -B / 2)-P WII h WII -P α h α

[0157] =47000(2-1.43)-1.3×4.5(3.5-2)-0.1785×3.5-7.0945

[0158] =9.44525t·m>0

[0159] When overturning edge AD:

[0160]

[0161] Therefore, the dynamic stability requirements are met.

[0162] 3. Anti-storm stability

[0163] Working condition III, when a storm strikes during non-working hours:

[0164] (A) The boom is perpendicular to the trolley track and faces the sea side.

[0165] Fixed part PG1 =2000kg, center of gravity coordinates (0,0);

[0166] Rotating part P G2 =45000kg, center of gravity coordinates (1.49, 0);

[0167] The machine weighs 47,000 kg, and the center of gravity is 45,000*1.49 / 47,000=1.43;

[0168] Capsizing edge BC, wind direction (along the boom from sea to land) or capsizing edge AB, wind direction (wind perpendicular to the boom);

[0169] When the overturning edge BC, ∑M=P G (S / 2+X G )-1.2P wIII ·h wIII (P1000 south of Shanghai, 1800 at sea) = 47(3+1.43)-1.2×1.4×1×1.8×3.5×5.1

[0170] =154.2316t·m>0

[0171] When the overturned edge AB, ∑M=P G (B / 2+y G )-1.2P' wIII ·h' wIII

[0172] =47(2-0)-1.2×1.4×1×1.8×3.5×6.863

[0173] =21.362t·m>0

[0174] (B) The boom is parallel to the trolley track.

[0175] Fixed part P G1 =2000kg, center of gravity coordinates (0,0);

[0176] Rotating part P G2 =45000kg, center of gravity coordinates (0, 1.49);

[0177] The machine weighs 47000kg, the center of gravity is 0, 45000*1.49 / 47000=1.43;

[0178] When the overturned edge AB, ∑M=P G (B / 2+y G )-1.2P wIII ·h wIII

[0179] =47(2+1.43)-1.2×1.4×1.8×3.5×5.1

[0180] =107.232t·m>0

[0181] When overturning edge AD,

[0182] Therefore, the stability requirements when non-working maximum wind load is applied are met.

[0183] 4. Anti-tilt stability

[0184] Working condition IV, with backward working wind load, assuming sudden unloading in the air, the unloading payload is 4.5t;

[0185] (A) The boom is perpendicular to the trolley track and faces the sea side.

[0186] The overturning edge BC, wind direction (along the boom from sea to land), and overturning moment are:

[0187] ∑M=P G (S / 2+x G )-0.2P I (R max +t+S / 2)-Pw II hw II

[0188] =47(3+1.43)-0.2×4.5(3.5+1.085+3)-0.1785×3.5

[0189] =200.76t·m>0

[0190] (B) The arm is parallel to the trolley track.

[0191] The machine weighs 47000kg, the center of gravity is 0, 45000*1.49 / 47000=1.43;

[0192] When the overturning edge is CD, the wind is from front to back along the boom;

[0193] ∑M=P G (B / 2+y G )-0.2P I (R max +t+B / 2)-P WII h WII

[0194] =47(2+1.43)-0.2×4.5(3.5+1.085+2)-1.785×3.5

[0195] =149.036t·m>0

[0196] (C) AD side capsizes, wind from land to sea,

[0197] ∑M=P G (S / 2-x G )-0.2P I (R max -t+B / 2)-P WII h WII

[0198] =47(3-1.43)-0.2×4.5(3.5-1.085+3)-1.785×3.5

[0199] =62.67t·m>0

[0200] Therefore, the stability requirements during sudden unloading are met, and in summary, the overall stability requirements are met.

[0201] Drive unit selection

[0202] 1. Wire rope

[0203] According to the maximum static tension and the working level of the mechanism, select the wire rope model, determine the wire rope diameter, and the total length of the wire rope.

[0204] The maximum static pressure of the wire rope wound into the drum (unit: KN) is:

[0205]

[0206] Wherein, m is the number of wire rope branches1;

[0207] The lifting device is a hook. The lifting mechanism is composed of two identical motor drive systems to realize the hook operation. Therefore, only one set of transmission system components needs to be determined. In a single set of drive devices, each drum bears 55% of the load, that is,

[0208] η Σ is the total transmission efficiency of the wire rope system, and the calculation formula is

[0209] Where η1 is the efficiency of a single pulley, which is 0.98 for rolling bearings; η D is the efficiency of the guide pulley. In this design, there are 4 guide pulleys, so η D =0.98 4 =0.92237; a is the wire rope pulley ratio, and from the structure we know that a=1;

[0210] The minimum breaking force of the wire rope is (unit: KN): F0 = nS max=5.6×22.335=125.076;

[0211] Where n is the safety factor of the wire rope. According to the working level M6, the safety factor of the moving steel core wire rope is n = 5.6.

[0212] Determine the minimum diameter of the wire rope according to the conditions (unit: mm):

[0213] Select the wire rope model: 166×19S-IWR 1770B. Its nominal breaking force is 161kN and the wire rope diameter is d = 16mm.

[0214] Select No. 4 wire rope pressure plate.

[0215] 2. Pulley

[0216] Determine the minimum pulley diameter based on the wire rope diameter and the working level of the mechanism, and select the appropriate pulley from the sample.

[0217] When the wire rope diameter d = 16mm, the pulley diameter D = h·d = 22.4×16 = 358.4mm, the pulley rope groove bottom radius R≈(0.53~0.60)d=9mm, and the rope groove wall angle 2β≈35°~45° should be selected.

[0218] Pulley diameter D0 = hd = 22.4 × 16 ≈ 360 mm. Here, h is the sheave-rope diameter ratio coefficient. Based on the M6 ​​operating class and the gantry crane, we use 22.4. D0 = D + d = 344 + 16 = 360 mm, meaning the pulley groove bottom diameter is 344 mm. Model: Pulley B1 16 × 360 × -100JB / 9005.5-1999. Use BT 100JB / T9005.5-1999 spacer rings, with dimensions of 100 mm inner diameter, 110 mm outer diameter, and 6 mm thickness.

[0219] 3. Reel

[0220] (1) Winding diameter of the reel (unit: mm): D 1，min =h1d=20×16=320, where according to the working level of the mechanism M6, look up Table 5-4 and get: reel h1=20, according to Table 5-5 common reel diameters, the initial reel diameter D1=600mm.

[0221] From the calculation of the roll geometric dimensions in Table 5-6, we can get:

[0222] Rope groove radius R = (0.53 ~ 0.56) d = 8.5mm;

[0223] Rope groove depth (standard groove) H1 = (0.25 ~ 0.4) d = 5mm;

[0224] Rope groove pitch (standard groove) t1 = d + 3 = 19 mm;

[0225] Coil thickness (steel coil) δ≈d=16mm;

[0226] (2) Winding length

[0227] Winding length for single-layer winding of single pulley block:

[0228] L d =L0+2L1+L2=190+2×38+2×16=298mm;

[0229] In the formula: L1 is the length of the position occupied by the fixed wire rope head, generally L1 = (2 ~ 3) t = 2 × 19 = 38; L2 is the length of the free part at both ends of the drum, generally 1 to 2 turns; L0 is the length of the spiral part of the drum.

[0230] L0=((H max a) / (D1π)+Z1)t=((15×1) / (0.6×3.14)+2)×19≈190mm

[0231] Where H max is the total lifting height, H max =5+10=15m; Z1 is the additional safety circle number, take Z1=2;

[0232] Considering the design, processing allowance and allowable deflection angle, the actual roll length L = 500mm.

[0233] (3) Wall thickness of the roll

[0234] The thickness of the roll (steel roll) is δ≈d=16mm.

[0235] 4. Verification of the roll

[0236] (1) Force analysis of the reel

[0237] according to Figure 6 It can be seen that the maximum working tension of the wire rope acting on the drum (unit: KN): S max =γ'φ2S=1.2×1.44×22.335=38.595.

[0238] Where, γ' is the load increase coefficient, according to the lifting mechanism level M6, γ'=1.20 is taken from Table 2-24 of the textbook; φ2 is the lifting dynamic load coefficient, according to the lifting state level HC2;

[0239]

[0240] Maximum bending moment (unit: kN·m):

[0241] Torque (unit: kN·m): T = S max D1=38.595×0.6=23.157;

[0242] (2) Strength verification

[0243] The bottom diameter of the rope groove of the drum is D = 600-16 = 584mm. Since L / D = 500 / 582 < 3, it is necessary to check the maximum stress σ on the inner surface of the drum wall. y .

[0244] Mpa<[σ y ];

[0245] Table 5-7, A1 winding layer number 1, take 1, A2 take 0.75;

[0246] The roll material is Q345B, then its allowable compressive stress [σ y ]=σ s(屈服强度) / 1.5=230Mpa, so the calculated strength meets the requirements.

[0247] (3) Stability verification

[0248] When D>1200mm and L>2D, the drum may become locally unstable. Therefore, there is no need to perform stability verification on the drum.

[0249] Select the drum diameter D = 600mm, length L = 500mm, drum groove bottom radius R = 8.5mm, groove pitch t = 19mm, lifting height 15m, and magnification 1;

[0250] Roll 600×500-8.5×19-15×1JB / T90062-1999.

[0251] 5. Design of reel shaft and bearing

[0252] (1) Material of the reel shaft

[0253] The reel shaft is made of 45 quenched and tempered steel;

[0254] (2) Preliminary determination of the shaft diameter d based on torsional strength

[0255]

[0256] Where A is the coefficient. Since the material of the shaft is 45 quenched and tempered steel, 119 is taken from the table.

[0257] When designing a keyway on a shaft, if d>100mm, increase the calculated diameter by 3% to 5%, so:

[0258] d min ≥159.87×(1 + 5%) = 167.87mm, rounded up to 170mm.

[0259] (3) Drum bearing

[0260] The bearing is initially selected as 23034, with basic dimensions d = 170mm, D = 260mm, B = 67mm;

[0261] The equivalent dynamic load P is: P = f p (XF r + YF a ) = 1.8×(1×2 + 2.6×38.55) = 184.014kN, F a / F r <e = 0.26, X = 1, Y = 2.6, the maximum axial force caused by the drum wire rope is (p121) 38.595 * sin3° = 2kN, and the radial force is cos3°S max = 38.55kN.

[0262] Basic rated dynamic load:

[0263]

[0264] In the formula, the temperature coefficient f r = 1.0, expected life, looking up the table for cranes [L h = 18000 - 25000, take 19000, life index ε = 10 / 3 (roller bearing). The basic rated dynamic load of bearing 23034 is 445kN, so the bearing is suitable.

[0265] VI. Check the strength of the shaft according to the combined bending and torsion strength

[0266] For a general steel shaft, the equivalent stress of the dangerous section can be obtained by using the third strength theory: In the formula, M is the bending moment borne by the drum shaft, T is the torque borne by the drum shaft, and W is the moment of inertia of the cross-section with respect to the shaft,

[0267] So In the formula, [σ -1 = 300, for 45 quenched and tempered steel it is 300Mpa.

[0268] The check passes.

[0269] VII. Hoisting motor

[0270] (1) Calculation of steady-state power

[0271] The steady-state power of the motor (unit: KW) can be calculated by the following formula:

[0272]

[0273] Where: η is the total efficiency of a motor drive system, η = η Σ =η 筒 η 减 =0.94×0.97; m is the number of motors.

[0274] (2) Selection of motor

[0275] According to the table, the preliminary selection of the benchmark working system using the equivalent connection duration method is S3, JC=40%, CZ=300 times / hour, and G=G2=0.8.

[0276] The motor model selected is YZP250M-1 (variable frequency speed regulation three-phase asynchronous motor).

[0277] 8. Verification of hoisting motor

[0278] The expression is:

[0279] H is the coefficient, and for variable frequency asynchronous motors, H=2.2; λ m The maximum torque multiple is 2.8; P N =22.59, overload verification passed.

[0280] The heat calculation of the motor is expressed as: Then P≥P S , so the motor heating verification is passed.

[0281] 9. Motor of running mechanism

[0282] 1. Steady-state operating power of the trolley running mechanism motor

[0283]

[0284] P j =P m +P α +P wI =4495.43+0+107.1*9.81=5546.1N;

[0285] (1) Wheel rolling friction resistance

[0286] P m1 =PLf K *2 / D=P Σ *2f K / D

[0287] P ΣThe gravity of all the weight of the moving parts, including the hanging weight 47000*9.81N;

[0288] f K is the rolling friction arm of the wheel along the track. According to the table, the wheel material is steel, flat top, and the tread diameter is 500, so take 0.5;

[0289] D is the wheel tread diameter 500mm;

[0290] P m1 =PLf K *2 / D=P Σ *2f K / D=47000*9.81*2*0.5 / 500=922.14N;

[0291] (2) Friction resistance in wheel bearings

[0292] P m2 =P Σ μd / 2*2 / D=P Σ μd / D=47000*9.81*0.015*150 / 500=2074.815N;

[0293] μ is the bearing friction coefficient, which is 0.015 for ball or roller type;

[0294] d is the shaft diameter, which is 150 mm;

[0295] (3) Rail wheel rim friction is difficult to calculate, so the amplification factor C is used. f Consider its impact;

[0296] (4) Friction resistance of rail running, the above sum

[0297]

[0298] Look up the table, C f Take 1.5;

[0299] P a =m a (m G +m i )g=(m G +m i )gtanα=xtan0=0,m i is the mass of the crane, α is the slope coefficient;

[0300] 0.6P wII =0.1785*0.6=0.1071t

[0301] 2. Selection of motor for trolley running mechanism

[0302] Steady-state operating power of the trolley running mechanism motor:

[0303]

[0304] P j =P m +P α +P wI =4304.14+0+107.1*9.81=5354.79N

[0305] (1) Rolling friction resistance of rail wheels

[0306] P m1 =PLf K *2 / D=P Σ *2f K / D

[0307] P Σ The gravity of all the weight of the moving parts, including the hanging weight 47000*9.81N;

[0308] f K is the rolling friction arm of the wheel along the track. According to the table, the wheel material is steel, flat top, and the tread diameter is 500, so take 0.5;

[0309] D is the wheel tread diameter 500mm;

[0310] P m1 =PLf K *2 / D=P Σ *2f K / D=45000*9.81*2*0.5 / 500=882.9N;

[0311] (2) Friction resistance in wheel bearings

[0312] P m2 =P Σ μd / 2*2 / D=P Σ μd / D

[0313] μ is the bearing friction coefficient, which is 0.015 for ball or roller type;

[0314] d is the shaft diameter, which is 150 mm;

[0315] Pm2=P Σ μd / 2*2 / D=P Σ μd / D=45000*9.81*0.015*150 / 500=1986.525N

[0316] (3) Rail wheel rim friction is difficult to calculate, so the amplification factor C is used. f Consider its impact;

[0317] (4) Friction resistance of rail running, add the above:

[0318]

[0319] C f Take 1.5;

[0320] P a =m a (m G +m i )g=(m G +m i )gtanα=xtan0=0(micrane mass,αslope coefficient);

[0321] 0.6PwII=0.1785*0.6=0.1071t.

[0322] Sea wave compensation hydraulic system

[0323] 1. Basic Principles

[0324] The active compensation hydraulic system is based on the integrated application of sensor technology, control algorithms and hydraulic actuators. The system monitors the operating status of the hydraulic system in real time through sensors and transmits the monitoring data to the controller. The controller accurately controls the hydraulic actuators according to the preset control algorithm to achieve active adjustment of the hydraulic system parameters. The crane wave compensation device is equipped with four hydraulic cylinders. If the ship is impacted by waves and causes the crane to deviate to a certain angle, four independent hydraulic cylinders are required to restore the original balanced position. That is, these four hydraulic cylinders do not interfere with each other. Figure 7 .

[0325] 2. Control Design and Simulation of Wave Compensation Device

[0326] A. System composition

[0327] 1) Torsion motor: used to provide appropriate speed and torque for the hydraulic pump;

[0328] 2) Oil cylinder: provide sufficient hydraulic oil;

[0329] 3) Overflow valve: used as a safety valve to control the maximum pressure of the oil in the system;

[0330] 4) Directional valve: changes the direction of oil flow, thereby controlling the direction of the hydraulic actuator;

[0331] 5) Hydraulic cylinder: As an actuator, the oil pushes the piston in the hydraulic cylinder to move and thus push the load to move;

[0332] 6) Filter: filter out impurities in the oil;

[0333] 7) Displacement sensor: detects the displacement of the actuator in real time;

[0334] B. System Simulation

[0335] According to the system's workload and working environment, select the appropriate hydraulic cylinder to ensure that the load capacity of the hydraulic cylinder meets the system requirements. Here, select a hydraulic cylinder that takes the mass of the hydraulic rod into account.

[0336] Motor speed: 2000 rpm, overflow valve opening pressure: 30MPa, directional valve opening maximum current: 40mA, directional valve maximum flow: 5L / min, hydraulic pump rodless cavity diameter: 25mm, hydraulic cylinder rod diameter: 12mm, hydraulic cylinder maximum extension length: 3m.

[0337] Figure 8 To independently control the simulation principle diagram of four hydraulic cylinders, the input signal of the main valve makes the main valve open, and the control signals of the other four directional valves are as follows Figure 9 shown.

[0338] Each directional valve is first input with a negative potential, then a positive potential, and finally a negative potential. From the simulation principle diagram, we can see that the negative potential causes the hydraulic cylinder piston rod to extend, and the positive potential causes the hydraulic cylinder piston rod to retract. The final displacement simulation diagram of the hydraulic cylinder end is Figure 10 ;Depend on Figure 10 It can be seen that the speed is the same during the rising process of the hydraulic cylinders, and the speed is also the same when the four hydraulic cylinders rise simultaneously, indicating that the pressure and flow of the system are sufficient and independent of each other.

[0339] C. Hydraulic cylinder position control

[0340] like Figure 11 In order to make the hydraulic rod reach a specific position, it is necessary to add a displacement sensor to the end of the hydraulic rod, and adding PID control to the system can realize the position control of the hydraulic cylinder.

[0341] Depend on Figure 12 It can be seen that the target position is set for each hydraulic cylinder. The simulation results show that all four hydraulic cylinders can reach the target position before 4 seconds. The rising time of the system is relatively long. The PID parameters will be adjusted later to shorten the time required to reach the target position.

[0342] D. Offset control of crane actuators

[0343] like Figure 13Considering the force of waves on the ship, which causes the ship to deflect at a certain angle, and ultimately the crane actuator to deflect at a certain angle, to maintain the cargo in a horizontal position, it is necessary to adjust the extension and retraction of the four hydraulic cylinders to restore the cargo to its original horizontal position. This article only discusses the case where the crane claw deflects at a certain angle around a certain axis.

[0344] Assume that the crane claw is offset counterclockwise by a certain angle due to the action of waves. In order to restore the cargo to a horizontal position, hydraulic cylinders 1 and 3 need to be extended, while hydraulic cylinders 2 and 4 need to be retracted. The distances moved by the four hydraulic cylinders are:

[0345]

[0346] Based on this principle, the simulation schematic diagram is designed as follows Figure 14 .

[0347] The initial extension length of the hydraulic cylinder piston is set to 2.5m, and the counterclockwise offset angle of the crane claw is set to 15 degrees. The PID parameters are adjusted to reduce the time required for the system to reach the target position. Finally, the image of the displacement of the four hydraulic rods changing with time is obtained as follows: Figure 15 .

[0348] The simulation results show that hydraulic cylinders 1 and 3 rise, hydraulic cylinders 2 and 4 fall, and by adjusting the PID parameters, the system can reach a stable state before 0.5s.

[0349] E. Wave heave compensation

[0350] Ocean wave signals are highly complex, random, and irregular, making their motion difficult to accurately describe using simple mathematical models. To simulate the random nature of ocean waves, a method often employed is the superposition of multiple simple harmonic waves of varying frequencies and phases, known as the Longuet-Higgins wave superposition model. In the Longuet-Higgins model, the irregular ocean wave signal ζ(t) can be represented as the superposition of a series of simple harmonic waves, as shown in the following formula:

[0351]

[0352] For the ζ(t) function at a fixed point at sea level, its general expression is to set both ξ and η to 0 and ignore higher harmonics, then the formula can be simplified to the following formula:

[0353]

[0354] Among them, ε i is the initial phase of the i-th harmonic, which is a random variable and is uniformly distributed in the range of [0, 2π]; ω i Represents the angular frequency, through the expression It can be obtained, where ω m =0.2rad / s,ω M =2.5rad / s;ζ ai is the amplitude of the i-th harmonic.

[0355] To further describe the spectral characteristics of ocean wave signals, the Pierson-Moskowitz (PM) spectrum is often used. The PM spectrum is a spectral model derived from sea state parameters such as wind speed and wave height, and is widely used to simulate irregular waves. The expression of the PM spectrum is as follows:

[0356]

[0357] T1 is the characteristic period, unit is s; S ζ (ω i ) is the angular frequency ω i The corresponding spectral density, unit is m 2 ·s; amplitude ζ ai It can be calculated by the following formula:

[0358]

[0359] In practice, the “spectral density” S is often used ζ To describe the ocean waves, it represents the ocean energy per unit frequency. ai It can be obtained from the spectral density.

[0360] Using the sea state parameters of Level 4, we simulated a binary irregular wave, taking a significant wave height of 1.85m and an average wave period of 6.7s. Substitute these into the formula and sum them (taking 24 wave components).

[0361] The final generated wave signal is as follows Figure 16 shown.

[0362] According to the literature, when the compensation speed is equal to the heave speed of the ship, the motion state of the load remains unchanged, thus ensuring normal operation. Based on this principle, it can be concluded that when the secondary wave signal is input, the heave compensation simulation principle diagram of the crane is as follows: Figure 17 shown.

[0363] The simulation results of the hydraulic cylinder displacement are shown in the figure Figure 18 As shown, since all four hydraulic cylinders move simultaneously, only the displacement of one is required. When the waves rise, the cargo must descend to maintain its original position, verifying the compensation principle described above. The simulation diagram shows that after compensation, the maximum displacement of the cargo relative to the ground is 0.035m.

[0364] In the description of this application, it should be understood that the terms "up", "down", "front", "back", "left", "right", "vertical", "horizontal", "top", "bottom", "inside", "outside", etc., indicating the orientation or position relationship, are based on the orientation or position relationship shown in the accompanying drawings, and are only for the convenience of describing this application and simplifying the description, and do not indicate or imply that the device or element referred to must have a specific orientation, be constructed and operated in a specific orientation, and therefore cannot be understood as a limitation on this application.

[0365] Those skilled in the art will appreciate that, in addition to implementing the system, device, and various modules provided by the present invention in purely computer-readable program code, it is entirely possible to implement the same program in the form of logic gates, switches, application-specific integrated circuits, programmable logic controllers, embedded microcontrollers, and the like by logically programming the method steps. Therefore, the system, device, and various modules provided by the present invention can be considered a hardware component, and the modules included therein for implementing various programs can also be considered structures within the hardware component; the modules for implementing various functions can also be considered both software programs for implementing the method and structures within the hardware component.

[0366] The above describes specific embodiments of the present invention. It should be understood that the present invention is not limited to the specific embodiments described above, and those skilled in the art may make various changes or modifications within the scope of the claims, which do not affect the essence of the present invention. The embodiments of this application and the features in the embodiments may be combined with each other in any manner unless there is a conflict.

Claims

1. A ship-to-shore dual-purpose crane for level 4 waves, characterized by: include: The track base is fixed to the foundation with bolts, and its gauge and length are adapted to the water level changes during the flood season and the dry season; The trolley travel base adopts a triangular truss structure, which is stably supported by oblique steel frames and horizontal support beams. The upper part of the base is provided with a trolley travel track. The trolley travel base is equipped with a 360° rotating structure to support and move the arm and achieve horizontal and vertical movement; The boom is used to support the grabbing mechanism and can be extended and adjusted in angle according to operational needs. Its extension range is calculated based on the moment balance formula to adapt to the height difference of ±10 meters between the ship and the shore; The lifting height of the grabbing mechanism is divided into upper and lower sections. The total lifting height covers the handling requirements under different water levels, and the dynamic load range is calculated using the lifting dynamic load coefficient formula; The wave compensation device independently controls the displacement of four hydraulic cylinders to effectively compensate for four degrees of freedom: heave, pitch, roll, and yaw. It provides heave compensation and ship roll compensation capabilities, and cooperates with the movement of the trolley to compensate for six degrees of freedom, thereby maintaining the stability of the crane. The trolley walking base is connected to the track base by bolts, the trolley walking base is installed on the upper track of the trolley walking base, the arm is hinged to the trolley walking base through a rotating structure, the grabbing mechanism is suspended at the end of the arm, and the hydraulic cylinder of the wave compensation device is symmetrically arranged at the connection between the trolley walking base and the arm.

2. The ship-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The triangular truss structure of the trolley walking base meets the following design conditions: The axial load of the truss members is optimized based on the material strength formula. Calculate the cross-sectional area, where σ is the allowable stress of the material, F is the axial force, and A is the cross-sectional area of ​​the rod; The self-weight of the truss is calculated by the formula of equivalent section moment of inertia Calculation to achieve lightweight design, where I i is the section moment of inertia of a single member, L i is the length of a single rod, L total is the total length of the truss; The base stiffness is calculated through the node displacement equation Verification is performed to ensure that the deformation under sea state level 4 does not exceed the safety threshold, where δ is the node displacement, E is the material elastic modulus, I is the section moment of inertia, and L is the length of the rod.

3. The ship-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The 360° rotation structure of the trolley walking base is achieved by the following methods: The selection of slewing bearings is based on the calculation of the support reaction force distribution, which meets the following requirements: where R max is the maximum bearing reaction force, M 回转 is the rotational torque, D 轴承 is the bearing diameter; The power of the slewing drive motor is given by the formula Determine, where T is the rotational resistance torque, ω is the angular velocity, and η is the transmission efficiency.

4. The ship-to-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The telescopic and angle adjustment of the boom is controlled by: The telescopic length of the boom is based on the moment balance formula F 伸缩 ·L 臂架 =G 负载 ·H 高度差 , adapted to the ±10m height difference between ship and shore, where F 伸缩 is the boom extension driving force, L 臂架 is the effective length of the boom, G 负载 is the load weight of the grabbing mechanism, H 高度差 is the maximum height difference between ship and shore; Boom inclination angle through hydraulic cylinder thrust formula Adjust to ensure the positioning accuracy of the gripping mechanism under dynamic load, where F 液压 is the hydraulic cylinder output thrust, M 负载 L is the moment of load on the hinge point of the boom, 力臂 is the length of the hydraulic cylinder force arm.

5. The ship-to-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The lifting dynamic load coefficient of the grabbing mechanism is calculated by the following formula: in, is the minimum dynamic load coefficient, β2 is the speed influence factor, v q is the lifting speed.

6. The ship-to-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The hydraulic cylinder control method of the wave compensation device includes: According to the ship's heave displacement signal, the hydraulic cylinder displacement command is generated through the PID control algorithm, where the error integral term where K p , K i , K d are the proportional, integral and differential coefficients respectively, e(t) is the displacement error signal, and e(τ) is the ship heave displacement signal; The heave compensation is calculated by the acceleration integral formula Calculate, the compensation error is less than the set threshold, where α(τ′) is the ship heave acceleration; Pitch and roll compensation is achieved through the balance equation ∑F of the hydraulic cylinder thrust and the boom inertia moment 液压 r = I α realization, where F 液压 is the output thrust of the hydraulic cylinder, r is the lever arm, I is the moment of inertia of the arm, and α is the angular acceleration.

7. The ship-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The overall stability of the crane is verified by the following steps: Basic stability: Based on the geometric relationship diagram of the support reaction force, establish the center of gravity coordinate equation Solve the distribution of support reaction force under the critical state of overturning edge, where G i is the weight of each component, x i is the coordinate of the center of gravity of each component; Dynamic stability: Combined with the overturning moment formula M under wind load 风 =P·A·h, verifying the critical values ​​of wind speed and wind pressure, where P is wind pressure, A is windward area, and h is the effective height; Anti-storm stability: through the moment balance equation M under the maximum wind pressure condition in the non-working state 抗倾 ≥M 风 Calculate the anti-overturning coefficient, where M 抗倾 The anti-overturning moment generated by the crane's own weight.

8. The ship-to-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The steel wire rope of the drive device is selected to meet the following requirements: The breaking force of the wire rope is calculated by the formula F 破断 =n·F 工作 ·η 滑轮 Calculate, where n is the safety factor, F 工作 is the maximum working tension of the wire rope, η 滑轮 is the pulley efficiency; The pulley diameter is determined by the sheave diameter ratio formula D = h·d, where h is the sheave diameter ratio coefficient and d is the wire rope diameter.

9. The ship-to-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The friction resistance of the running mechanism is calculated as follows: Trolley running friction resistance: comprehensive rolling friction resistance Bearing friction resistance and slope resistance F 坡道 =G·sinα to derive the total resistance, where μ 滚动 、μ 轴承 are the rolling friction coefficient and the bearing friction coefficient respectively, α is the angular acceleration, D is the pulley diameter, G is the mass of the running mechanism, and d is the wire rope diameter; Frictional resistance of the trolley: Based on the weight distribution formula F 摩擦 =μ·G and dynamic load magnification factor C f Correction, where μ is the comprehensive friction coefficient.

10. The ship-to-shore dual-purpose crane for level 4 waves according to claim 1, characterized in that: The simulation verification of the wave compensation device includes: Generate random signals based on the Longuet-Higgins wave superposition model, Simulate the fourth-order sea state, where ζai is the amplitude of the i-th harmonic, n is the total number of harmonics, ωi is the angular frequency, and ∈i is the random initial phase; Calculate the amplitude via the Pierson-Moskowitz spectrum Verify that the cargo displacement after compensation is Δs≤0.035m, where S ζ (ω i ) is the spectral density; Energy efficiency optimization is achieved through the hydraulic pump speed adjustment formula Q = k·n and the power recovery rate of the energy recovery device Implementation, where Q is the flow rate, k is the pump displacement coefficient, n is the pump speed, P 回收 、P 输入 For recovered and input power.

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