Table of Contents

1 Design original data        1

 2 Working conditions        1

 3 Overall scheme        1

 4 Stacker lifting mechanism design        2

 4.1 Hoist main components        2

 4.2 Motor power Calculation        3

 4.3 Shaft design        4

 4.4 Chain drive main parameters selection        5

 5 Stacker walking mechanism design        6

 5.1 Overview        6

 5.2 Determine wheel diameter        7

 5.3 Select track        8

 5.4 Select gear motor        9

 5.5 Select wheel spindle        11

 6 Stacker fork design        11

 6.1 Fork drive overall selection 6.2        11

 Transmission device movement and dynamic parameters calculation        12

 6.3 Fork belt drive design calculation        13

 6.4 Fork drive gear, rack calculation        14

 7 Three-dimensional warehouse shelf design        18

 7.1 Shelf overview        18

 7.2 Design parameters        18

 7.3 Static pressure analysis        18

 8 Summary        20

Design of single column roadway stacker

1 Design raw data

2 Working conditions

1. Working conditions: two-shift work (each shift is calculated as 8h)
2. Working environment: indoor, less dust
3. Service life: depreciation period of 8 years, overhaul once every 4 years;
4. Manufacturing conditions and batch size: ordinary medium. small manufacturer

3 Overall program

The single column roadway stacker is mainly composed of lifting mechanism, walking mechanism, fork and shelf, which completes the automatic transportation and storage of items and realizes the efficient operation mode of unmanned warehouse. This article designs the single column laneway stacker mainly for the smallcargo (maximum weight does notexceed 1000N) but the design. The specific design drawing is as follows:

Design of lifting mechanism of 4 stacker

4.1 Main components of hoist

The power of the motor drives the chain wheel to run through the coupling device, thus driving the chain to lift the goods. Because the main shaft of the hoist is rotated, which drives the operation of the sprocket, and drives the pallet to rise and fall through the chain, so in the design process, the requirements for the shaft are very high. Here, a separate motor is used to control the lifting mechanism, and in order to prevent the lifting platform from falling down during power failure, causing damage or safety accidents, the brake motor with brake function is selected here. Therefore, the main components are: motor, sliding cylinder, chain, sprocket, column. The following figure is the overall structure diagram:

1. Chain hoisting mode  

Chain lifting mode is the use of chain sprocket transmission, so as to enhance the goods.

The chain lifting method requires that the kinetic energy is first transferred to the top of the column by the transmission chain, and then the cargo is lifted by the lifting chain. The chain drive is driven by the meshing of the sprocket teeth with the chain links. The chain drive has no elastic sliding and overall slipping phenomenon, which can ensure accurate transmission ratio and high transmission efficiency. The chain flexible component does not need to be installed too tightly, so the pressure on the shaft is not large, the overall size of the chain drive is small, the structure is compact, and it can work under high temperature and humid conditions. Compared with the wire rope, the chain has the advantage of large flexibility, the number of sprocket teeth can be small, so the diameter is small, the structure is compact, and its disadvantage is that the sensitivity to impact is large, the possibility of sudden breakage is large, and the wear is also faster. In addition, it cannot be used for high speed, and the lifting chain is used for machinery with small lifting capacity, small lifting height and low lifting speed.

2. Take-off and landing sliding cylinder structure:

  The take-off and landing sliding cylinder structure is mainly composed of sliding cylinder, guide wheel group and other parts. The take-off and landing sliding cylinder structure is mainly used to fix the cargo platform and bear the weight of the cargo platform and the cargo. The design is to meet the requirements of strength and stiffness. The loading platform module comprises a plurality of guide wheel sets which mainly comprise guide wheels, bearings and guide wheel shafts. The guide wheel group is mainly in contact with the upright post or the guide rail, so that the loading platform module is connected with the upright post module. The following figure shows the structure of take-off and landing slide cylinder:

4.2 Motor power calculation

Usually we choose the motor first to consider whether the motor can provide the torque and speed required by the load. The power required by the hoisting motor can be calculated according to the static power of the hoisting mechanism under full load and stable operation, which is calculated as follows:

where   P-maximum lifting load (N);

v-rated lifting speed (m/min);

Total efficiency of the lifting mechanism.

Total weight of lifting platform and cargo: sliding cylinder, V=0.2m/s,

In consideration of other uncertain factors, the braking motor of lifting motor model YEJ 90L-4 is taken from the table,=1.5kw, n=1440r/min

    Select the motor of the hoisting mechanism according to the static power required by the mechanism. After selecting the motor, the overload capacity and heat generation of the motor shall be checked. The hoisting mechanism requires that the motor can be hoisted when there is voltage loss. 25 times the rated lifting capacity. Therefore, the power of the motor should meet the requirements of the formula to ensure sufficient overload capacity.

Where, N---rated power of motor (kw) ;

        P---Maximum lifting load (N);

        v---Rated lifting speed (m/min);

        ---Total efficiency of lifting mechanism.

        H---coefficient, H =2 for asynchronous motor. 1, DC motor takes H =1. 4;

        Z---the number of driving motors of the mechanism.

It meets the requirements after verification.

4.3 shaft design

The shaft is an important part of the machine, its function is to support the rotating parts, so that the rotating parts have a certain working position, and transfer motion and power. In general, reasonable structure and sufficient strength are the basic requirements that the design of the shaft must meet. If the structural design of the shaft is unreasonable, it will affect the processing and assembly of the shaft, increase the manufacturing cost, andeven affect the strength and stiffness of the shaft. If the strength of the shaft is insufficient,plastic deformation or fracture failure will occur,making it unable to work properly. The general design steps for the shaft are:

Select materials according to job requirements.  

② Estimate the basic diameter of the shaft.  

Structural design of shaft.  

1. Estimate the basic diameter of the shaft

In the design of the shaft,the minimum diameter of the twisted shaft section on the shaft is generally estimated according to the torque transmitted by the shaft, and the structural design of the shaft is carried out with it as the basic reference size.

In this topic, the material of the main shaft is 45 steel, through the table can be []= 60 MP,= 103 45 钢，通过查表可得[]=60 MP， =103

P=×η=1.5×0.8=1.2kw

It is also known that n=n1=501440=28.8r/min

∴==35.7mm

The required d shall be the diameter of the shaft section subjected to torsion,i.e. the shaft diameter at the belt driving wheel. Considering the influence of other factors,the standard diameter d=40mm is taken.  

2.Structural design of shaft

(1)Initial determination of the diameter of each shaft section

Considering the coupling connection and other factors, calculate and check the mechanical manual to determine the diameter of each shaft end as follows:

Diameter of each shaft end

(2) Determine the length of each shaft section

(3)Circumferential fixing of transmission parts, sprocket and belt transmission are A-type common flat keys, sprocket is 16×9 GB1096.

(4)For other sizes, refer to the installation dimensions of 6209 C-type bearing (see bearing manual) for processing convenience. The radius of the transition filet on the shaft is r=1mm, and the chamber of the shaft end is 2×45.

4.4 Selection of main parameters of chain drive

1. Selection of transmission ratio

When the transmission ratio is too large, it will lead to the small sprocket wrap angle is too small, and the number of meshing teeth is reduced, which will accelerate the wear of the sprocket teeth and easily lead to the phenomenon of tooth skipping, so the wrap angle is preferably not less than 120 °. For this reason, the transmission ratio is limited to ≤7. In order to ensure that the pallet moves vertically along the column, the driving and driven sprockets have the same size, so the transmission ratio i = 1.

2. Number of sprocket teeth

In this design, since the chain speed is 0.2m/s, it belongs to low speed. The number of teeth Z1 of the small sprocket can be selected according to the chain speed reference table.

The number of teeth should be determined according to the transmission ratio i, and Z1= Z2=17 is taken from the table.

3. Determine the calculated power

where    P-Transmitted power, KW;

K A-working condition coefficient, refer to table to know K A=1.3;

K Z-coefficient of number of teeth of driving sprocket, K Z=0.85 according to table;

KP-Multi-row chain coefficient, single row chain KP =1.

4. Select chain type and pitch

If the transmission ratio of the reducer is i=31.5, then

calculated power

The chain pitch can be checked to select the chain number of 20A, then the chain pitch is p=31.75mm

5. Check chain speed and determine lubrication method

v< 6m/s, in the low speed range, meet the conditions.

According To v=0.26 m/s and chain number 20A, the lubrication mode is determined to be periodic manual lubrication mode.

6. chain wheel

the steel 45 is selected for quenching and temper (40 - 50) HRC

The design of the traveling mechanism of 5 - 5 stacker

5.1 Overview

Speed of travel, i.e. dead weight

According to the working requirements of the traveling mechanism, the utilization level of the mechanism is T7, and the load condition of the mechanism is L2, so the working type of the mechanism is determined to be M7.

5.2 Determining the wheel diameter

Take, take the wheel speed.

        Before loading, the force on the two wheels of the stacker is not uniform. Set the non-uniform coefficient as 1.4 and take the driving wheel for calculation.

Take, you can get:

Take the wheel.

5.3 select track

track size drawing

According to the wheel diameter size, determine the track type is 12 and its size specifications.

5.4 Select gear motor

        Add resistance factor to account for other running resistances

Therefore, select the gear motor to view the SEW gear motor for selection. Considering overload, select the gear motor SEW S57 DV100M4 P=2.2KW T=295N.m

Gear motor parameter diagram

size drawing of gear motor

Matching diagram of motor and reducer

5.5 Wheel spindle

        The minimum diameter of the wheel spindle is calculated to be 30 mm.

6 stacker fork design

6.1 Overall selection of fork drive

Fork is mainly used to complete the loading of goods, using the telescopic plate of the fork itself to complete the removal of goods from the shelf, or put the goods on the shelf.

1) Design parameters

2) Selection of transmission scheme

Forks mainly complete the role of loading goods, the force is large, so the need for multi-stage deceleration, first by the motor through the reducer for deceleration, and then through the pulley for secondary deceleration, transfer torque on the gear rack. Because the shelf plate needs to be telescopic, the double rack and pinion mechanism is used to complete the telescopic function.

3) Motor selection,

By friction F f =μ*m*g=0.85*100*9.8=833N.

Check the table to know:

η ν band =0.99 η tooth =0.98 .9

The overall efficiency of the transmission is then:

η total = η with η teeth η teeth η teeth = 0.99 x 0.98 x 0.98 x 0.98 = 0.88总=η带η齿η齿η齿=0.99X0.98X0.98X0.98=0.88

Static power of motor operation: P 0 =F f *V=210W

Therefore, the selected motor model is: DC brushless motor 80BL89S25-430TK0 power P=250W, rotation speed n m =3000r/min

4) Determine the transmission ratio

Select drive gear wheel diameter D=50mm

==76.4r/min

iTotal ===39.26

Take i minus =25, i.e. iband =1.57

5) Determine the reducer

The model of reducer is PX80N025SW, the allowable torque of low-speed shaft is 70N*m, and the reduction ratio is 25

6.2 Calculation of kinematic and dynamic parameters of transmission device

1) Calculation of rotating speed of each shaft:

n

n==120r/min

n=76.4r/min

2) Calculation of input power of each shaft

P0=Pηm=250X0.9=225W

P 1 =P 0 η minus= 225 x 0.85 = 191 W

P2 = P1 η band = 191 x 0.99 = 189 W

P3 = P2nHet = 189 x 0.98 = 185.22 W

P4 = P3nHet = 185 x 0.98 =181.5W

3)Calculation of input torque of each shaft

T0=9550P0/n0=716.25N·mm

T1=9550P1/n1=15200.4N·mm

T2=9550P2/n2=23625N·mm

6.3 Design and calculation of fork belt drive

1) Incoming power P1=191W, and working condition coefficient Ka =1.7

Design power: P d=Ka *P 1 =1.7X191=324.7W

2) Determine the type and pitch of the synchronous belt: from the design power: P d =324.7W, n 1 =120r/min s,the selected belt type is H and the pitch is 12.7 mm

3)Select the number of teeth of the small pulley:

Check the table to obtain: the number of teeth of the small belt pulley:Z1 =14,the number of teeth of the large belt pulley:Z2= Z1*i=14X1.57=21.98Take the number of standard teeth: Z2 =22

4) Determine Pitch diameter: d1= =56.6mm d2= =89.0mm

5) Determine the pitch line length of the synchronous belt: L p

initial center distance

Therefore, 101.92mm< α0< 291.2mm is initially determined to be 200mm

=4.64Lp=2a*cos++ =629.9mm

Take the standard pitch line length closest to the calculated valueL p=635mm

6) Calculate the number of synchronous belt teeth: =50

7) Calculation of transmissioncenter distance α

a = where

α = 202.55mm, α = 203mm=202.55mm ,取α =203mm

8) Determine the synchronous bandwidth:

8.1 Determination of reference power rating of H-type synchronous belt

In which: allowable working pressure: 2100.85N

      m: mass per unitlength: 0.448

      V==0.36m/s

Substitute:

8.2) Calculate the number of meshing teeth of small pulley: Zm =

8.3)Determine the actual bandwidth: bPreliminary selection

P meets design requirements

In summary, the pulley parameters are as follows:

Parameters of H-type synchronous belt

6.4 Calculation of Fork Transmission Gear and Rack

It is known that the input power P2=185.22W, the rotation speed of the pinion n1 =76.4r/min, and the gear ratio u1=1 due to the gear rack transmission. Driven by motor, with service life of 8 years (300 days per year) and 2-shift system, then:

Select gear type, accuracy class, material and number of teeth

1.1) According to the transmission scheme, the straight gear rack transmission is selected. b. The fork is a general working machine with low speed, so the 8-level precision is selected.

1.2 Material selection. According to the table, the material of pinion is 40Cr (quenched and tempered), the hardness is 280HBS, the rack is 45 steel (quenched and tempered), the hardness is 240HBS, and the hardness difference between the two materials is 40HBS.

1.3) Primary pinion tooth number Z 1 =19

Design according to contact strength of tooth surface

1.4) to determine each calculated value in the formula

1.4.1) Trial load factor Kt=1.3

1.4.2) Calculate the torque transmitted by the pinion T =23625N·mm.

1.4.2) Tooth width coefficient: According to Table 10-7 of Mechanical Design textbook P205, it is obtained: = 1

f. Elastic influence coefficient of material: According to Table 10-6 of P201 in Mechanical Design textbook,=189.8MP 1/2

1.4.3) According to Figure 10-21 on page P209 of the textbook Mechanical Design, the contact fatigue strength of the pinion is Mpa according to the hardness of the tooth surface, and the contact fatigue strength of the rack is 550Mpa.

1.4.4)Calculate the number of stress cycles: Calculate the number of stress cycles from Formula 10-13 in the textbook P206 of Mechanical Design

=60×74.6×1×（2×8×300×8） =1.76×108

1.4.5) Contact fatigue life coefficient: refer to Figure 10-19 of Mechanical Design textbook P207 to obtain: K HN1 =0.95        

1.4.6) Allowable contact fatigue stress of gear: Taking the failure probability as 1%, the safety factor S=1, and applying the formula 10-12 of P 205 in the textbook of Mechanical Design, we get:

       []==570

1.4.7) Calculation:

Pitch circle diameter d of pinion

wherein the gear rack transmission ratio

d 1t =34.9mm

1.5)Calculate the peripheral velocity v  

1.6). Calculate tooth width b and module m nt

b=

mnt=

1.7). Calculate the ratio of tooth width to tooth height

     h=2.25 m =2.25×1.84=4.14mm

=

1.8). Calculated load factor K

Refer to Table 10-2 of Mechanical Design Textbook P 193 for coefficient of use K A =1.25. According to v=0.14m/s and Class 8 precision, refer to Figure 10-8 of P 194 for coefficient of dynamic load K V =1.12, refer to Table 10-4 of Mechanical Design Textbook P 197 for coefficient K Hβ =1.42, refer to Table 10-3 of Mechanical Design Textbook P 195 for coefficient K Hα =1.

Therefore, the load factor:

K=

1.9)Calibrate the calculated reference circle diameter according to the actual load factor

d1=

1.10)Calculated modulus m

m =

2)Design according to bending strength of tooth root

≥

Determine the calculated values in each publication

2.1)Load factor K

K=

2.2) Primary selection of tooth width coefficient: according to symmetrical arrangement, it is found from the table that =1

2.3). Check tooth profile coefficient Y Fa and stress correction coefficient Y Sa

Check the mechanical design textbook P 200 table 10-5:

Tooth form factor Y Fa1 =2.85

Stress correction factor Y Sa1 =1.54 Sa2 =1.785

2.4)Calculation of pinion

Refer to Table 10-20 in textbook P208 to obtain the bending fatigue strength limit:  

pinion

Refer to Table 10-18 of P206 in the textbook to obtain the bending fatigue life coefficient:

KFN1=0.92      

Taking the bending fatigue safety factor S=1.4, then

2.5)design calculation        

Comparing the calculation results, the modulus m calculated from the contact fatigue strength of the tooth surface is larger than that calculated from the bending fatigue strength of the tooth root. Taking m=2.5mm, the bending strength can be satisfied. However, in order to meet the contact fatigue strength at the same time, it is necessary to calculate the number of teeth according to the reference circle diameter d1=40.22 calculated by the contact fatigue strength. So by:

Z1=, Undercutting will occur. When the number of teeth isadjusted to 19 teeth, i.e. d=47.5mm

Circumferential speed: V= is similar to the set speed V=0.2 m/s, which meets the requirements.

3)geometric size calculation

3.1) Pinion geometry calculation:

According to tooth width coefficient b=47.5mm

addendum height:

Root height: =3.125mm

addendum circle:

Root circle diameter:

3.2)Calculation of geometric parameters of rack:

Root height: =3.125mm

addendum height:

Rack width:

     Rack height: H= h is the height from the bottom of the gear to the bottom of the rack

     Pitch: P = mm：P=mm

     Tooth thickness: S=

     Slot width: e=

Center distance: L= mm

7. Design of cargo rack for stereoscopic warehouse

7.1 Rack Overview

Warehouse automation construction is a systematic project, and the three-dimensional shelf plays a role in storing goods in the whole system.

and has motion matching with the hoisting machine and the stacking machine.

7.2 Design parameters

The three-dimensional warehouse is 8 columns and 7 floors

Total height is 3.5m

8 shelves per layer

size: L * W * H =550*500*550mm

7.3 static pressure analysis

The role of the shelf is to carry the load, so it must have a certain stiffness and strength. Therefore, the static pressure finite element analysis of the most important stress parts in a single cargo compartment is carried out with solidworks as follows:

1) Example Attributes

Conclusions

The stress deformation is within the acceptable range and can meet the product requirements.

8 Summary

In this course design, the team members through the observation and analysis of the existing stacker structure and functional principles, in the self-developed parameters, the structure of the stacker was re-designed. In the course design of this project, on the one hand, the team members have the equipment to measure, on the other hand, by comparing the various programs, analyzing the advantages and disadvantages of the program, and then designing the optimal program of this work.

In the process of parameter formulation, the course design team members consult a large number of documents, including the Mechanical Design Manual, as well as the industry standards of the stacker for parameter formulation and part selection, and strive to design the program in line with relevant standards.

In the process of design and calculation, we also refer to the design and calculation process of predecessors, strictly according to the formula for calculation, forsome difficult to calculate the function, also use matlab software for calculation.

Worksin the process of three-dimensional modeling, using solidowrks for drawing, and the key parts of the force using finite element analysis software for analysis, from the results that the design meets the requirements.

Finally, I would like to thank the team members for their unity and ability to answer each other's questions, so that they can complete a high-quality course design within a limited time. Finally, I would like to thank Mr. Xiao and the laboratory teachers for their guidance.