Process Engineering Design Pressure Vessel - Essay Example

The expected urge pressure from of the 15 ATM storage pressure is 16. 5 ATM. This will be taken as the operating pressure throughout the design. As the actual pressure inside the vessel will be is 0. 979 from SOURCE. The total volume of the daily quantity of methane at the specified conditions is calculated below. However the vessels must not fall below O gauge pressure inside, there must be at least one ATM absolute pressure inside each vessel otherwise they will fail due to the difference in pressure. So total amount of storage needs to account for this.

We Will Write A Custom Essay Sample On
ANY TOPIC SPECIFICALLY
FOR YOU

For Only $13.90/page


order now

Assuming a linear relationship exists between pressure and illume at constant temperature 1 ATM would correlate to one fifteenth of the volume, such that total volume required. Allowances for extra volume in-case of train delays etc will be made later in the design. CHOICE OF END CAP To avoid designing a shell and then having to make adjustments such that it fit a specific end cap the end cap was chosen first and the shell designed around it. Seem-Ellipsoidal end caps were chosen as the operating pressure exceeds 15 bar (pop).

Hot pressed caps will be used to save on costs. Whilst a greater range and larger sizes are available in the hot spun method of manufacturing it is not worth the higher cost. The largest available semi ellipsoidal available by hot pressed manufacturing from Australia Pressure Vessel Heads catalogue has an internal diameter of 2731 mm. This is approaching the largest size that can be easily transported by conventional means. It will need to be manufactured to order but this is unavoidable as the largest stock size is gorgon smaller.

If the vessel falls into class 3 the joint efficiency, , will be 0. 7, as specified by table 1. 6 AS 1210 (pig 23). Applying the minimum shell thickness equation at the design pressure: So as a class 3 vessel the minimum thickness of the shell wall is 26. 78 mm. This fails the criteria outlined by table 1. 7, AS 1210 (pig 23) as Note that as well, justifying choice of an AY material. An AY grade material would have provided no benefit. If it is class B, still using a double butt weld, the weld efficiency is increased to 0. 8.

Re applying the minimum thickness equation yields For the vessels to fall under class three the diameter would need to be more than halved. This would result in either very long, uneconomic vessels or many smaller vessels. In either scenario it is not worth the savings made from leaner design restrictions and safety checks. A corrosive substance to steel, so a minimal corrosion scenario is assumed. Assume 0. 1 mm of corrosion per year (Collusion and Richardson V). Designing or a tank life of 20 years gives a corrosion allowance of 2 mm.

Total shell wall thickness given by mm However steel sheets only available in either 25 mm or 28 mm. 28 mm would be overburdening for no real benefit, so 25 mm sheet will is to be used. This reduces the corrosion allowance to 1. 6 mm, which is still acceptable. This is still below 32 mm, so vessel is still classed as B. Cap Thickness Firstly the same material as the shell will be used for the end caps, so the tensile strength will be the same, mamma. Table 1. 6 AS 1210 (pig 24) shows the joint efficiency for double butt welds for circular welds is 0. . Minimum highness for a semi-ellipsoidal end cap is given by K is a factor dependent on the ratio between the diameter of the cap and its height. The height of the cap is given in Australian Pressure Vessel Heads data sheet and for the selected cap is given as 682 mm. The diameter, D, used is the internal diameter of the shell and end cap. And applying the minimum end cap thickness equation gives Internal corrosion for the end cap will be the same as that for the shell, so apply corrosion allowance of 1. 6 mm.

However from the Australian Pressure Vessel Heads data sheet it can be seen it is only available in 25 mm or 32 mm thickness. The 25 mm thick variant will be used. It will decrease the corrosion factor to 1. 57 mm, however this is still acceptable. And to prevent buckling (pig 35, ASIA 210) is fine. And as the shell and end caps have same internal diameter and thickness a strong weld will be easy. Shell Aspect Ratio and Volume The diameter was determined by the end cap as 2731 mm. Initially let This is the middle of the recommended range for pressurized vessels (page 27) and close to the optimal of 3.

This gives a shell length of 9558. 50 mm. However steel sheets only available in mm increments and cutting to size old be wasteful and provide no benefits. Better to just have slightly larger tanks. So adjust shell length, L, to 9600 mm. Gives a new aspect ratio of This is still within the recommended range. Volume Total volume of pressure vessel given by Volume of madcap is specified in data sheet as 2661 liters and the volume of the shell is easily calculated. As previously shown of tank space is required to store a days methane.

So to calculate the required number of tanks To allow for downtime, maintenance and late trains 20 tanks shall be used. Inlet Nozzle The pressure vessel requires an inlet nozzle, outlet nozzle, pressure relief Alva and drain pipe. Table 3 from AS 4343 (pig 9) specifies that extremely flammable gases are rated as Very Harmful Gas (VBG). Methane falls into this category. From table 1. 4 AS 4041 (pig 8) unless design pressure exceeds 2 Amp class 3 piping is acceptable. Class 3 piping acceptable. To calculate the size of inlet nozzle required the volumetric floodwater of methane is to be calculated.

Note that for the following calculations to make literal sense a constant methane pressure of 1. 67 Amp ( must be assumed through the inlet nozzle, ii the driving force from the pump. Per day, therefore the volumetric floodwater, Q, is A target gas velocity of is chosen as it sits in the middle of the optimum range. Using this velocity the cross sectional area of the pipe, , can be calculated And the diameter can easily be calculated from the area Note that this is the internal diameter. PIPE MATERIAL BBS 806 BBS 3601 430 ERE was chosen for all piping.

Design conditions are neither unusually high or low, and this choice of material reflects that. F-room Table 02 AS 4041 (pig 59) it can be seen that at the design temperature this material exhibits a tensile strength, f, of 183 Amp. From table 3. 12. 3 AS 4041 the design factor, M, for class 3 piping is 0. . As it is piping seamless welds are to be used which have a weld joint factor, e, Of 1 . The pressure design minimum pipe thickness, , can be calculated using the internal diameter based equation The contribution of pressure to the necessary design thickness is very small.

The total minimum pipe thickness is a function of the pressure design term and the corrosion allowance can be calculated by As the material is a carbon alloy and methane being relatively UN-corrosive a corrosion allowance Of 2 mm will be used. Rearranging the formulae yields And so the theoretical outer diameter So a theoretical inner diameter of 46. 0 mm and outer diameter of 49. 13 mm. DEN 40 schedule 40 (Fluoroscope Pipe catalogue, page 21), with an outer diameter of 48. 3 mm and a thickness of 3. 68 mm, fits the theoretical requirements the closest.