TABLE OF CONTENTS

TABLE OF CONTENTS……………………………………………….…………1

TABLE OF CONTENTS……………………………………………….…………2

CHAPTER 1: CALCULATE TRANSMISSION SYSTEMS……………..….…3

I. SELECT AN ELECTRIC MOTOR…………………………………………..….3

II. RECALCULATE TRANSMISSION RATIO DISTTRIBUTION………..….…4

III. OTHER PARAMETERS…………………………………………………..…..5

CHAPTER 2: ROLLING CHAIN DESIGN………………………………...……6

I. CHOOSING CHAIN TYPE:…………………………………………….………6

II. DESIGN PROCESS:………………………………………………………...…8

III. STRENGTH CHECKING…………………………………………...…………9

IV.CHAIN DRIVE PARAMETER…………………………………………...……9

CHAPTER 3: DESIGNING GEARS TRANSMISSION……………….…...…11

I. DESIGN PROCESS……………………………………………………….……11

1. Choosing the material for all gears:……………………………………………11

2. Calculation for the first helical gear:………………………………………..…13

3. Strength checking for first pair of gears:…………………………………......16

4. Calculation for the straight gear:…………………………………………...….18

5. Strength checking for first pair of gears:……………………………………….21

II. CHECKING LUBRICATION CONDITIONS IN OIL…………………...…….23

CHAPTER 4: DESIGNING SHAFTS AND KEYS………………………...……24

I. INITIAL PARAMETERS…………………………..……………………….…..24

II. DESIGN PROCES:……………………………………………………………24

CHAPTER 5: SELECTING COUPLING………………………………..……..38

I. INITIAL PARAMETERS……………………………..…………………...…….38

II. DESIGN PROCESS…………………………………………………….……..38

CHAPTER 6: SELECTING BEARINGS………………………………..……...39

I. SELECTING BEARING TYPE…………………………………………..……..39

II. SELECTING BEARING GRADE22……………………………………..……39

III. SELECTING BEARING PARAMETERS…………………………….…..….39

CHAPTER 7: DESIGNING GEAR BOX COVER AND FINALIZING……….43

I. DESIGNING THE GEAR BOX COVER……………………………………...43

II LUBRICATION FOR THE COMPONENTS…………………………...…….44

REFERENCES…………………………………………………………...……….46

CHAPTER 1: CALCULATE TRANSMISSION SYSTEMS

I. SELECT AN ELECTRIC MOTOR

The maximum power of working element:

Where:      

F: force on conveyer, F = 7000 N

v: belt velocity, v = 1 m/s

=> P_lv = 7 (kW)

The average power of the working element:

Based on the given Load diagram to determine the average power of the working element.

The efficiencies of gears and chain and bearing are chosen from table 2.3 p.19[1], given that the gears and chain are covered, it is suggested that choosing the value in a specific range; given no practical experiences, all the middle values from those ranges are chosen for better accuracy.

=> u = 0,8857

Number of revolutions of the motor:

Choose motor according to p.21[3]:

=> Choose motor 4A132M4Y3 with P = 11 kW, n = 1458 rpm

IV. RECALCULATE TRANSMISSION RATIO DISTTRIBUTION

This value satisfies the suggestion that u_chain = 2 5, according to table 2.4p.21[1]

V. OTHER PARAMETERS

Maximum power on shafts: (Equation p.49[1])

Note: The power of the motor’s axis after attached to the system is smaller than its initial state, due to friction lost from a coupling and a pair of bearing, however this lost is small so we can neglect it.

Table of properties: (using maximum power and maximum torque)

CHAPTER 2: ROLLING CHAIN DESIGN

I. CHOOSING CHAIN TYPE:

Due to low level load, low chain velocity, choose simple block chain, the center line of 2 sprocket are on the horizontal plane (α = 0 ) for the system to be simple. Also, single row chain is chosen for the rolling chain system design.

II. DESIGN PROCESS:

Calculate the number of teeth for driving sprocket

Based on table 5.4 p.80[1] with  = 2; It is suggested that choosing from 27…25 teeth; choose the middle value for better precision:

Calculated chain transmission power (Equation 5.3/81):

To satisfy the condition ≤ [P]

Choose [P] = 19.3 kW from table 5.5 p.81[1]

Based on table 5.5 p.81[1], choose rpm as recommended to be closest to the chain’s driving sprocket value (which is 122 rpm), allowable power [P] = 19.3 (kW), choose:

Pin diameter: 9.55 mm

Roller diameter: 19.05 mm

Choose the basic axis distance based on equation 5.11 p.84[1]:

It is suggested that choosing a = (30 … 50)p; choose the middle value for better precision:

VII. STRENGTH CHECKING

Where:

Q = 88.5 kN: maximum load (chosen from the table 5.2 p.78[1] given our pitch  diameter = 31.75 mm):

Chose C45 quenched with hardness HB210 has the maximum stress of  = 600 MPa (from table 6.1 p.92[1]) so the chain drive satisfies the strength conditions

VIII. CHAIN DRIVE PARAMETER

Table of properties:

Type of chain: Single row

Power: P (kW): 18.7426

Roller length: B (mm): 19.05

Pin length: b (mm): 46

Maximum load: Q (N): 88500

...........

Number of teeth: z (teeth): 25 ; 50

CHAPTER 3: DESIGNING GEARS TRANSMISSION

I. DESIGN PROCESS

1. Choosing the material for all gears:

According to table 6.1 p.92[1], without any specific requirement, choose widely used material type I material (HB ≤ 350), to facilitate the gear’s teeth cutting procedure, make it more accurate, lessen the need for stronger material for shafts, but also provide high allowable stress;  ≥  + (10 … 15)HB so as to reduce the friction between gears :

For the driving gear we choose steel 45 quenched with rigidity ; ultimate strength  = 850 MPa; yield limit :  580 MPa

For the driven gear we choose steel 45 quenched with rigidity ; ultimate strength  = 750 MPa; yield limit : 450 MPa

Total working time for both gears:

Where,

Serving years: 5 years

Working hours per shift: 8

Total working days per year: 230 days/year

Total shifts per day: 2

2. Calculation for the first helical gear:

Gear module:

Based on suggestion not to choose m ≤ 1.5 to 2 mm which will break the gear teeth easily when overload, choose m = 2.5 mm

Driven gear teeth:

=> Choose 86 teeth

Geometry parameters of the gears: (Equations in table 6.11 p.104[1])

Driven gear reference diameter:

d_w2 = u x d_w1 = 4.05 x = 257.1963 mm

Driving gear speed:

According to table 6.13, 6.14 p.107[1], given that  = 4.8 m/s, choose the degree of accuracy is 8, choose the undistributed load coefficient K_Ha= 1.0869; K_Fa= 1.2262

3. Strength checking for first pair of gears:

Determine the bending stress for driving gear:

Where: the parameters are notated in the previous part

Overloaded coefficient:

 based on table P.13p.237[1]:

Pitch circle diameter of driven gear:

=> d₂ = 275 mm

Choose, (the shifting coefficients) based on table 6.9 p.100[1], because displacement is not needed that it might affect other qualities of gear, such as reducing messing efficiency.

Driven gear reference diameter:

d_w2 = u x d_w1 = 2.973 x =269.3878 mm

5. Strength checking for first pair of gears:

Overloaded coefficient:

Based on table P.13p.237[1]: k_qt = 2,2

II. CHECKING LUBRICATION CONDITIONS IN OIL

According to p.463[2] : h2 = 5.625

However, h₂ should be higher than 10mm. Therefore, the lowest oil level to cover the teeth of the gear 2 is chosen to be: h_min = 10mm

=> The oil lubrication conditions is satisfied.

CHAPTER 4: DESIGNING SHAFTS AND KEYS

I. INITIAL PARAMETERS

Choose shaft material: steel C45 standard quenched as it is suggested for gear boxes with average load [1]. The other properties of this material (taken from table 6.1 p.92[1]): Hardness: 250 HB

Calculate initial parameters for shaft II:

Driven helical gear hub length:

l_m22 = l_m24 =52,5 mm (10.10 p.189[1])

Width of bearing:

b₀₂ =21 mm (table10.2 p.189[1])

Elastic coupling length:

 k₃ + h_n + l_m14 + 0.5 x b₀₁ = 12 + 12 + 45 + 9.5 = 78.5 mm

Length of shaft III: (use table 10.4 p.191[1])

l_m33: Sprocket wheel hub length, l_m33 = 1.5 x d₃ = 1.5 x 50 = 60 mm

b_o3: Width of bearing, b_o3 = 27mm

k₃, h_n = 12, chosen above.

SELECTING KEY:

Choose flat key (rectangular spunk key) as it is simple and widely used.

Length of key at critical sections: (use formula as suggested in p.173[1])

Other parameters are chosen based on table 9.1a p.173[1] given the shaft diameter at the key position, presented in table 4.1.

Testing the key’s stress and shear strength:

=> All the keys satisfy the strength test

CHEKING SHAFT FATIGUE STRENGTH

Shaft’s diameters chosen by using moment formula in section III have not been through some condition analyses referring to fatigue stress such as period of stress changing, the concentrated stress, size coefficient, surface quality… That is why strength checking using the formula below is needed.  

Table 4.2: Parameters for strength check 1

All safety factors calculated in table 4.6 are greater than the allowable safety factor  (which is the middle chosen from the range suggested in p.195[1]).

=> The shafts satisfy fatigue condition.

TEST FOR STATIC DURABILITY

To prevent large elasticity deformation on shafts or break down when being subjected to sudden load (such as when starting the motor). 

CHAPTER 5: SELECTING COUPLING

I. INITIAL PARAMETERS

Torque on motor shaft:  = 53.9158 Nmm

Choose flexible coupling as it is widely used, simple, light and cheap, can reduce resonance, impacts when the load cycle changes.

The material for manufacturing is quenched C45 steel.

II. DESIGN PROCESS

These parameters are chosen from table 16-10a p.68[1] that satisfy the conditions: , , (of shaft I), given our shaft diameter, obtained from table P1.7 p.242[1].

These parameters are chosen from table 16-10b p.68[1] given that the torque T of the coupling chosen above.

=> The elastic ring satisfies the strength test.

CHAPTER 6: SELECTING BEARINGS

I. SELECTING BEARING TYPE

For shaft I, the highest ratio: , this satisfy the condition that angular contact bearing is suggested to be chosen: .

=> Select single row angular contact ball bearings with contact angle  as they can both support radial forces and axial forces, while also working with moderate speed with low noise.

=> Select single row angular contact ball bearings as they cheap and simple and widely used and no other specific requirements are needed.

II. SELECTING BEARING GRADE22

Choose grade 0 for reduction gearbox based on suggestion p.213[1]

III. SELECTING BEARING PARAMETERS

The parameters are selected based on dynamic load, to prevent wearing and corrosion, or on static load to prevent surplus deformation.

All the shafts rotate with the speed larger than 1 rpm ð Use dynamic load to choose the bearings.

With:  is coefficient of centripetal load (chosen from table 11.4 p.215[1], based on practical coefficient e; for shaft II and III, e  is chosen based on the equation i , with i=1 is the number of rows of the bearing; is the ability to sustain stable load, chosen from table P2.12 p.264[1]; for shaft I,

=> Choose the largest size bearing for shafts I,II,III (p.254,264 [1]) so they can both satisfy the fitness to the shaft.

CHAPTER 7: DESIGNING GEAR BOX COVER AND FINALIZING

I. DESIGNING THE GEAR BOX COVER

Use the material to make the cover is gray cast iron GX15-32 as it is mostly used, the gear box is made by casting as the method can provide great hardness and light weight.

The parameters for the gear box cover are chosen based on table 18.1 p.85[1] and table 18.2 p.88[1].

With:  is the largest center distance between 2 shafts

 is the length and width of the gear box respectively

Choose the distance between the bottom of the gear to the bottom surface of the gear box cover of the gear box cover  so that the debris at the bottom of the gear box won’t be splatter up, also to satisfy the condition of oil needed for heat disposing, which is about 3 liters for 11.5kW.

Choose the number of oil cap’s screw  (table 18.2 p.88[1])

II. LUBRICATION FOR THE COMPONENTS

Apply the method of soaking the gear box’s inner elements in oil, as it is common and suitable for gears with speed lower than 12 m/s, with the total amount of oil = 3 liters for 11.5kW transmission ratio.

The oil type to use is AK-20, which is suitable for gear box, is chosen from table 18.13 p.101[1] to satisfy the conditions in table 18.13 p.101[1].

For the bearing grease is use for lubrication, based on the working heat and the number of revolutions. LGMT2 grease is chosen as it is suitable for small and medium size bearings, with high water resistance as well as wear resistance.

REFERENCES

[1]. Trịnh Chất, Lê Văn Uyển. Tính toán thiết kế hệ thống dẫn động cơ khí tập 1,2. 2006

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