pbe_19961101_37.pdf - Powder and Bulk Engineering

Powder and Bulk Engineering, November 1996

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m Gary A. Chubb and Keith E. McGuire Peabody TecTank r_

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There’s a lot of information available on designing storage tanks. But how do you properly use the information to ensure your tank can resist the vertical and lateral loads exerted on it? This article discusses how to adapt portions of published codes to design a tank that handles your loading conditions without being expensively overdesigned or dangerously underdesigned. The first sections cover calculations for shell strength and vertical compression, lateral tension, and roof design. The final sections provide a tank design example and explain how to implement these design calculations.

fter calculating loads on your storage tank, the critical part of the design process is designing the A tank shell, roof, and base to resist these loads.

[Editor’s note: This article doesn’t cover calculating the loads. Find more information in published sources or con- tact a storage equipment consultant or supplier.]

Despite the many published codes for various tanks, no US standard design manual exists for tanks made of carbon steel, stainless steel, or common alloys. [Editor’s note: Throughout this article, the term code designates published specifications, codes, guidelines, and standards.] In fact, the most appropriate shell strength for this type of tank has never been clearly identified. The result is that you can misapply available code information, lead- ing to either a grossly overdesigned tank that’s exces- sively costly or an underdesigned tank that’s unsafe.

When writing a tank specification, perhaps you refer to a published tank design code, such as API 650 or ASME B96.1, found in the company library or another source.

But chances are you don’t know exactly how the code was intended and, very often, the code isn’t applicable to your tank design.

Why not? Most such codes are intended for tanks storing liquids or gases rather than bulk solids. Many codes don’t even cover the shell construction material you’re consider- ing. Some commonly misapplied codes are the following:

American Petroleum Institute (API) 65O’is for a welded tank “whose entire bottom is uniformly supported” and for a tank storing “petroleum, petroleum products, and other liquid products commonly handled by.. .the industry” as defined in the code’s “Scope” section - in other words, a flat-bottom oil storage tank.

APZ 6202 is also for a welded tank and, as defined in the code’s “Scope” section, is applicable “to tanks that are intended to (a) hold or store liquids with gases or vapors above their surface or (b) hold or store gases or vapors alone.” The API codes are intended for carbon steel tanks and generally don’t apply to stainless steel or alloy tanks.

American Society of Mechanical Engineers (ASME) B96.1 is “for welded aluminum-alloy,. . .flat bottom,. . .tanks storing liquids” for the chemical industry. It doesn’t address other construction materials or specifically conical bottoms or tank shell loads other than hydraulic (liquid) forces.

American Water Works Association (AWWA) D l 004and D1035are for welded and bolted water storage tanks. They don’t discuss construction materials other than carbon steel.

ASME Section VIZZDivision Z,6 also called the “pressure vessel code” or “the code,” is a conservative code for tanks and vessels with design pressures of 15 psi or

38 Powder and Bulk Engineering, November 1996

higher. It addresses many construction materials but doesn’t address specific stored products, although the code implies it’s for liquid and gas applications. The code’s high safety factors typically result in an overdesigned tank. The code also tends to yield an uneconomi- cal design for the typical bulk solids tank at atmospheric pressure.

A look at how loads are exerted differently on a liquid tank than on a bulk solids tank will help to explain how these codes can be misapplied to the latter. Generally a liquid exerts hoop pressure (lateral tension) on a cylinder’s walls based on the liquid’s specific gravity and depth. But the liquid doesn’t exert compressive loads (which are vertical) into the shell because there’s no fric- tional force from the liquid to the shell. However, wind and seismic loads, the dead load (the weight of the tank and any attached equipment), and the live load (the weight of items not always on the tank, such as snow) exert compressive loads into the liquid tank’s shell. In the previously described codes, the shell’s total allowable compressive load typically includes the shell-stiffening effects of the shell’s hoop pressure. A safety factor of 4.0 or higher is common in these codes.

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2. Conversely, bulk solids exert both a hoop force from lat-

from the product’s friction on the shell. If you use your product’s bulk density to calculate a specific gravity and apply liquid tank design equations for hoop pressures, you’ll get an overly conservative tank design with a thicker shell than necessary. The compressive loads in the tank’s product zone (the cylindrical area) are less than just the product’s dead weight because product friction trans-

lating these loads based on the product’s mass or funnel flow discharge pattern is a science in itself and isn’t cov-

compressive load and hoop pressure, along with the cal-

operational loads. r_

rn You can readily adapt many parts of these codes to de- E. signing a bulk solids tank. The codes have valuable information on conical hoppers (load transfer calculations, design, and fabrication) and roof designs. The following

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fers some of that load to the shell as hoop pressure. Calcu-

ered in this article. The tank shell must safely resist the

culated wind forces, seismic loads, snow loads, and other

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sections discuss information from the codes that’s appropriate for designing a bulk solids tank at atmospheric pressure.


Calculations for shell strength and vertical compression

If a tank shell is a perfect cylindrical shape with a uniform compressive load, the shell’s critical buckling stress can be written as:

ocr = 0.6E t/r (1)

where a,, is the critical buckling stress (in psi), E is the construction material’s modulus of elasticity (in psi), t is the shell thickness (in inches), and r is the shell radius (in inches).

But because a bulk solids tank is never a perfect cylindrical shape, you need to analyze the tank shell with methods based on research and testing. Most available methods for calculating the shell’s critical buckling stress are based on the classical Equation 1 and use correction factors based on testing, experience, and good design practices. Most of the following equations allow the critical buckling stress to extend beyond the construction material’s yield strength and well into the material’s plastic failure and elastic deformation range (where the material is stressed so much it can’t return to its original shape). Because your tank shell probably will have defects and experience nonuniform loadings, it’s good design practice to limit the shell’s critical buckling stress to your construction material’s yield strength.

The calculations. VonKarman’s equation’ calculates the critical buckling stress as:

ocr = 0.194E t/r (2)

where all terms are the same as those for Equation 1 and a 0.194 correction factor brings the results into a more practical range.

Donnell’s equation* is empirical and based on data from 100 test cylinders and other researchers’ results.

The hypothesis is based on plastic failure and elastic deformation, which is reflected by the yield strength’s inclusion as one component in this equation. Even with this component, larger values of t/r still produce a critical buckling stress value over the yield strength limit:

(0.6t/r) - [(r/t)x10-7] (1 + 0.0O4E/oy)

ocr = E x

where ay is the yield strength limit.

(3 )

The National Aeronautics and Space Administration (NASA)9 developed an equation, with further development by Baker,’Ofor designing cylindrical structures in aircraft and space vehicles. Similar to the VonKarman equation (Equation 2), it has a correction factor based on r/t ratios. This calculated correction factor allows for a larger critical buckling stress for t/r values greater than 0.0025 and, because of the progressive strength increase allowed for larger t/r values, can be very practical for designing a tank shell. The equation for critical buckling stress is:

ocr = y x 0.6 Eth

where y is 1 - 0.901 (1 - e+) and @ is (1/16)(r/t)l~

Roark‘s equationl’for critical buckling stress is:

where p is Poisson’s ratio.

(4)

(5 )

As different as Equations 1 and 5 appear, they give the same results. In Roark’s research, tests showed that the critical buckling stress developed is usually only 40 to 60 percent of the calculated theoretical value. Thus Roark suggests using these reduced values for the design critical buckling stress by rewriting the equation as:

ocr = 0.3Et/r

44

Plotting the calculations for diflerent shell construction materials. Figure 1 plots the results of Equations 1 through 6 for the tank construction material A-36 carbon steel, using 36,000 psi for ay and 29,600,000 psi for E. Figure 2 plots the results of the equations for 5052 H-32 aluminum, using 23,000 psi for ay and 10,200,000 psi for E. The plots show the variations for critical buckling stress versus t/r for each equation. The yield strength limit in each plot is a dashed horizontal line that indicates each construction material’s theoretical plastic failure and elastic deformation point. If your construction material’s yield strength will be reduced by welding (as it is for aluminum), use the lower published value for yield strength, thus limiting the critical buckling stress to allow for the tank shell’s reduced yield strength.


Each equation plotted in Figures 1 and 2 forms a straight line except Equation 4. This reflects the equation’s use of r/t in the correction factor. The Equation 4 line curves up- ward for larger values of t / ~ You can expect additional inherent shell strength for the larger t/r values because the shell will be more stable and minor shell defects won’t have as much effect.

The Figure I and 2 plots show that the critical buckling stress Equations 2 and 3 lead to very conservative results, yielding a thick-walled tank design. Equation 1, the classical buckling equation, and Equation 5 give extremely liberal results; you shouldn’t use them because both equations are based on a perfect geometric shape with no defects. Equation 4 will provide the best tank design because it’s based on detailed testing and yields a practical solution with an expected tank shell strength increase due to the higher t/r ratios. However, you can also use other equations to design your bulk solids tank if you properly test and document them.

Once you select a critical buckling stress equation, consider which safety factors are appropriate for your design. Your product will exert hoop pressure on the shell, which stabilizes the shell and justifies a safety factor from 2.0 to 3.0 in the product zone. If your tank shell is supported by a skirt, with a hopper suspended within the shell, the skirt area won’t be supported or stiffened by hoop pressure or stored product. Doors, conveyor openings, and pipe pene- trations in the skirt can further reduce its stiffness. In this case use a skirt safety factor from 3.0 to 4.0.

Figure 1

Critical buckling stress calculated by Equations 1 through 6 for A-36 carbon steel

70,000 I / / / I I I Eauations 1 and 5 / Eauation 4 // I

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To determine the tank (or product zone) safety factor when applying Equation 4 to designing a skirt-supported tank, consider the tank height H (in feet) times the tank diameter D (in feet). Use a tank safety factor of 2.0 for a tank with H X D < 1,000 and 2.5 for a tank with H X D 2 1,000. Equation 4 also yields the skirt safety factor 3.0.

Based on this information, Figure 3 compares the t/r values versus the allowable stress for A-36 carbon steel using Equation 4. The upper curve (with safety factor 2.0) plots the shell’s allowable stress; the lower curve (with safety factor 3.0) plots the skirt’s allowable stress. The additional curves plot the allowable stresses for tanks that have no skirts and are based on codes API 620, AWWA D103, and ASME Section VI11 Division I. These curves are shown only for comparison, not because the codes are recommended for designing a bulk solids tank. The additional curves clearly show that each code requires a higher safety factor, thus yielding a more conservative design, than Equation 4. Also notice that each of the three curves has a slope that reduces as it reaches the carbon steel’s yield strength, revealing an inherent yield strength limit.

Figure 4 presents the same information for 5052 H-32 aluminum for both welded and unwelded (bolted or other) tanks. Curves for both tank types are shown because, unlike with A-36 carbon steel, welding affects the aluminum’s temper and results in a lower yield strength,

shown here by the dashed horizontal lines intersecting the Equation 4 shell and skirt curves. (The yield strength for unwelded tanks is represented by each curve’s continua- tion beyond this limit.) The figure also shows the yield strength’s effects: that is, the slope of each Equation 4 curve reduces as it reaches the yield strength limit. The allowable stresses from the code ASME Section VI11 Divi- sion I are also plotted, only for comparison.

Because the other equations (2,3,5, and 6) are straight- line equations, they aren’t plotted in the figures. You can easily apply them, with appropriate safety factors, to Fig- ures 3 and 4 to compare results. The plots in the figures are only for comparison in this article; don’t use them for tank design because you can easily get more exact results by adapting any of the equations to a computer program.

So in Figure 3, the maximum allowable stress using Equation 4 is achieved at a t/r of about 0.0045. In this case, a 12-foot-diameter tank shell would have a %-inch thickness, a 20-foot diameter shell would be thicker than % inch, and a 30-foot-diameter shell would be nearly % inch thick. Each is a very stable shell with a high t/r value that would minimize effects from defects.

In Figure 4, the maximum allowable stress using Equa- tion 4 is achieved at a t/r of about 0.007. In this case, the 12-foot-diameter tank shell would be % inch thick, a 20-


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foot-diameter shell would be nearly X inch thick, and a 30-foot-diameter shell would be 1 % inch thick. Each of these is also very stable and would minimize effects from defects.

the hoop pressure when designing the tank. You can use the following equation to calculate the resulting tension load:

You can use Equation 4 as a practical design approach for T = pr (7) a bulk solids tank because NASA's required reliability provides a high confidence level for applying the equation. Using Equation 4 is an economical alternative to using the highly conservative national codes written specifically for liquid or gas tanks. However, if you apply appropriate safety factors, you can achieve a reasonable tank design using the other equations.

where Tis the shell tension (in pounds per inch), p is the total resisted pressure (in psi), and r is the shell radius (in inches).

Then compare the calculated tension to the strength of the tank joint - either welded or bolted. For a welded joint, apply an appropriate joint efficiency factor based on the extent of your X-ray examination of the joint. Joint efficiency factors are typically 0.7 for no X-ray, 0.85 for spot X-ray, and 1 .O for 100 percent X-ra~. '-~,~For a bolted joint, apply the strength requirements of ANSI SG-671120r an-

Calculations for lateral tension Product stored in a tank both is restrained by the shell and, in the form of hoop pressure, exerts lateral tension on the shell. If the tank is operating under any significant pressure above atmospheric, you should add that pressure to

Powder and Bulk Engineering, November 1996 49

other recognized standard for tank construction. Because the tank pressure is typically low and adds additional strength to the shell, you generally don’t need to do a combined stress analysis of compressive and hoop forces.

If your tank has a suspended conical hopper, the product load results in tension from hoop pressure TI and radial pressure T2, as shown in Figure 5. Because a cone is a modified cylinder with a converging radius you can use Equation 7 for a cylinder, substituting the radius normal to the cone for the equation’s shell radius and checking the cone for hoop pressure resistance. Check about 1 foot down the hopper slope for the hoop pressure resistance. You can typically reduce the thickness of a long hopper near its outlet because the hopper’s converging radius di- minishes the pressure toward the outlet.

Also check the joint design for resistance to radial pressure, which is the force in the direction of product flow, using the same joint efficiency factors previously described. See equations in References 2,3,6,10, and l l for converting the lateral pressure in the tank’s cylindrical product zone to loads on the conical hopper. Also consider any other loads, such as an axial load or bending mo- ment from discharge equipment, that could exert

additional stress on the hopper. To design a tank hopper that handles several loading conditions, see equations in References 10 and 11.

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Attaching the hopper to the tank shell produces an inward pull on the hopper Q, called circumferential or compressive force and expressed in pounds. For a tank operating at or near atmospheric pressure, the shell is usually thick enough to resist this force. And because the load is compressive, you don’t need to apply a joint efficiency factor. You can then find the minimum area required to resist the compressive load (called the compression area), using an allowable stress of 0.5 av, with the equation:


Q A, = - 0.5 oY

where A, is the compression area (in square inches), Q is compressive force (in pounds), and ay is the compression area’s yield strength.

The compression area is that part of the shell and hopper that resists the compressive loads exerted by the hopper attachment, as shown in Figure 5. References 1,2, and 6 give calculations that define the compression area in different ways, but any of the calculations will give a satis- factory tank design. In Reference 1, the effective compression area A (in square inches) can include a part of the shell within a distance 16 times the shell thickness both above and below the attachment point and 16 times the cone thickness. Reference 2 gives the shell’s compression areaA as within a distance of 0.6(rt)”* in each direction on the shell and down the cone’s slope. The radius r used in the calculation is the shell radius for the shell’s compression area and the radius normal to the cone for the cone’s compression area. Reference 6 allows the area A to be within a distance of (rt)” in each direction, again includ- ing part of the cone.

You’ll need to add more compression area to resist compressive loads if A, as determined by Equation 8 exceeds the calculated areaA of the tank shell and hopper. You can add more area to this compression region in the form of a flat-rolled plate, rolled structural shape, or built-up mem- ber so the compression areaA is at least equal to A,. Refer- ences 1 and 3, which provide similar calculations for shallow cones (tank roofs, discussed in the next section), only require meeting minimum compression area limits and don’t specify shapes or sections to resist the compressive loads. Thus meeting the area requirement as calcu-

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lated in Equation 8 will provide a good design for your d tank’s compression region. - - cn 0

Calculations for roof design -0 - L - E A bulk solids tank typically has a roof to protect the prod- s ca cal but can also be flat. It must resist snow and other

specified live loads, dead loads, vacuum and pressure, and other loads. 0

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uct from weather and pollutants. The roof is usually coni-

For a flat or shallow-cone (less than 10 degrees from horizontal) roof, use a flat plate with enough roof rafters to reduce the required roof thickness. You can use any standard flat plate stress equation to determine the roof sheet thickness. Design the rafters to support the combined loads using Reference 4,13, or 14 or another recog-

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2. You can design a conical (10 degrees from horizontal or greater) roof using Reference 1,2, 3, or 10. A self-sup- porting roof is preferable, but not always feasible or eco- nomically possible. If required, you can add roof rafters to help distribute the loads over the roof surface, reducing the roof’s required thickness. See Reference 3,4, or 5 for how to size the rafters. If the roof itself can’t support the tank’s attached equipment, design support beams to carry these loads using the calculations in Reference 13 or 14 or in another recognized standard.

Tank design example Using the example of a steel tank 10 feet in diameter and 30 feet tall and specified according to the Uniform Build- ing Code,’5 the total static load at the tank base is 11 1,571 pounds and the total dynamic load is 261,946 pounds. [Editor’s note: These numbers for the static and dynamic loads are from “Part I: Development of design loads” in the authors’ original paper, “Bulk storage tank design considerations,” Proceedings of the 1996 Powder & Bulk Solids Conference, May 1996, Rosemont, Ill. See the original paper for more information on designing a tank according to different codes.]

According to Reference 15, you can increase the allowable stress for dynamic loading conditions by one-third. Thus you need to make and compare two calculations. Use Equation 4 to calculate allowable shell stresses for an


A-36 carbon steel tank, where ay is 36,000 psi and E is 29,600,000 psi, assuming %, ?/16-, and %-inch plate is available for fabricating the tank.

For %inch plate:

1116 60/0.1251/2 B =

= 1.369

y = 1 - 0.901 (1 -

= 0.328

For %-inch plate:

1116 6010.1875~/2 o =

= 1.118

y = 1 - 0.901 (1 - e-1.118)

= 0.394

Because the loads are at the tank base, the suggested safety factor is 3.0 and the allowable stress for the dynamic loading condition will increase by one-third. Rather than calculate actual shell stress for each thickness, convert the base load to a load per inch of circumference and convert the shell stress to an allowable load per inch. Then you can compare the allowable load per inch to the actual load per inch. Compare the allowable load per inch for:

% x t Static loading condition = - 3

and:

(ocr x t x 1.33) 3

Dynamic loading condition =

This gives the results in Table I.

Calculate the actual shell load in pounds per inch by dividing the load by the tank circumference. Then the load at the tank base becomes:

For X-inch plate:

1116 @ = -

6010.25 ~2

= 0.968

y = 1 - 0.901 (1 - e-0.968)

= 0.441

1 I1 3 7 1 pounds 120 inches x K

= 296 lblin for the static condition

and:

26 1,946 pounds 120 inches x n

= 695 lblin for the dynamic condition

Then:

0.328 x 0.6 x 29,600,000 x 0.125 60 Ocr =

= 12,136 psi for 118-inch plate

0.394 x 0.6 x 29,600,000 x 0.1875 60 Ccr =


0.441 x 0.6 x 29,600,000 x 0.25 60 Ocr =


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Then the allowable stress for %-inch plate becomes:

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For the calculated allowable loads, %-inch plate will satisfy the static condition, but the dynamic condition requires %-inch plate. Thus the tank skirt should be %-inch plate. It’s also standard practice to make the entire skirt one thickness unless the skirt is extremely tall.

To check the shell’s resistance to the calculated total compressive load from product, wind, seismic, live, and dead loads in the product zone, make the same calculations (obtaining the actual shell load in pounds per inch by dividing the load by the tank circumference) at the base of each horizontal ring seam. The shell thickness that resists the compressive load is usually enough to resist the hoop pressure unless the tank diameter is large or the hoop pressure is extremely high. Thus calculate the hoop pressure in the product zone assuming the same shell thickness required to resist the compressive load. You’ll need to increase the shell thickness if the thickness calculated to resist the hoop pressure is greater than the thickness required to resist the compressive load.

15,120 psi x 0.125 inch = 1,890 lb/in

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s Because the actual stress is 105 lb/in, the calculation shows that the 0.125-inch shell thickness is adequate. ca

d You can use a similar procedure to check the hopper thickness. Assume the product loads on the hopper at a % point 1 foot down the hopper slope are 3 10 lbhn for T, and 130 lbhn for T2. Q

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Use the larger of the two values for T and divide by the al- rn cn.

lowable stress in tension to calculate the minimum design thickness. Because the allowable stress was calculated as 15,120 psi for A-36 carbon steel with ajoint efficiency of 0.7, you can calculate: 2.

T = -

o a For the example tank, assume the calculated hoop pressure from product is 1.75 psi (250 pounds per square foot)

ness to resist the total compressive load is 0.125 inch. 3 10 lbiin 15,120 psi

- at 18 feet of product depth. Also assume the shell thick- -

= 0.0205 inch minimum

Then you canuseEquation7 to calculate the shell tension from this hoop pressure:

Because it’s impractical to fabricate a hopper from 0.0205-inch-thick material, compare this value to the minimum thickness that’s reasonable for fabricating a hopper and use the thicker of the two values. For the example tank, the hopper thickness will be 0.125 inch.

Check the shell-hopper attachment point for resistance to compression using Equation 8, assuming a compressive load of 11,000 pounds. Calculate the area required to resist compression compression

T = 1.75 psi x 60 inches = 105 lb/in

The allowable stress for the shell material in tension (from hoop pressure) is generally accepted as 0.6 uJ,, so you can calculate the shell’s allowable stress as:

as:

oa = 0.6 oY = 21,600psi

11,000 pounds A, = 0.5 x 36,000 psi

= 0.611 square inch

where a, is the allowable stress. If you assume the compression area is 16 times the shell thickness in each direction from the shell-hopper attachment point, calculate b e area as: If the shell seams will be welded without X-ray examina-

tion (for which the joint efficiency factor is 0.7), calculate the allowable stress as:

0, = 0.6 oYx 0.7 = 15,120psi

A = [2 (16~0.1875 inch) x 0.1875 inch] + [(16xO.125 inch) x 0.125 inch]

= 1.125 + 0.1406

= 1.27 1 square inches

0

‘I 0 73 -. ra 5 5. AWWA 0103: AWWA Standard for Factory-Coated Bolted Steel d

Tanks for Water Storage, American Water Works Association, 1987. 0 cn 0 -0

E 7. T. VonKarman and H.S. Shen, “The buckling of thin cylindrical

shells under axial compression,” Journal of Aero Science, No. 8, s 1941. 3 ca

-0 8. L.H. Donnell, “A new theory for buckling of thin cylinders under 0 axial compression and bending, Transactions of ASME, Vol. 56, d 1934. Q

%

6. ASME Section VIlI Division I: ASME Boiler & Pressure Vessel Code, The Amencan Society of Mechanical Engineers, July 1995.

S

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P, 3 Q m r_

9. “Buckling of thin-walled circular cylinders,” NASA SP8007: Space Vehicle Design Criteria Structures, August 1968.

10. E.H. Baker, L. Kovalevsky, and EL. Rish, Structural Analysis of 7 i

3

3

Shells, McGraw Hill, 1972. rn E.

edition, McGraw Hill, 1989. (D 11. Warren C. Young, Roark’s Formulas for Stress and Strain, 6th

2. 3 ra 12. AIS1 SG-671: Specifications for the Design of Cold-Fomted Steel

Structural Members, American Iron and Steel Institute.

56

Because the actual compression area exceeds the required area, the shell-hopper attachment point can adequately resist the hopper attachment’s compressive load without additional reinforcement.


How to implement these design calculations To design your bulk solids tank, work with a design engineer who has experience designing cylindrical shells and applying specific loads such as wind, seismic forces, snow, product, and internal pressure and vacuum. The engineer can help you design a high-quality tank based on practical calculations such as those in this article.

While the national codes listed in this article aren’t intended specifically for bulk solids tanks, they contain valuable design and fabrication information. Besides information on tolerances and quality assurance, the codes explain how to: reinforce the shell for nozzles or other items that penetrate the tank shell (such as fill lines, vent lines, and level indicators), design a roof, transfer loads to a conical hopper, anchor the tank, and more. An experienced design engineer knows how liquid tanks covered in the codes react differently than bulk solids tanks because of the differences between liquids and solids. By applying the loads and appropriate yield strength, the engineer can design your tank to provide economical, reliable bulk solids storage for many years.

Working with an experienced engineer can help you avoid designing a tank with overly high safety factors and high minimum shell and hopper thicknesses. Also make sure you build the tank on a solid, well-designed foundation and use quality fabrication standards. At minimum, apply reasonable tolerances for out-of-roundness measurements and shell alignment, require mill test certificates for the tank’s base metal, and rely on your fabrication shop’s or the supplier’s proven fabrication experience, certified welders, and qualified welding procedures. PBE

References 1. API-650: Welded Steel Tanks for Oil Storage, 9th edition, American

Petroleum Institute, July 1993.

2. API-620: Design and Construction of Large, Welded, Low-Pressure Storage Tanks, 8th edition, American Petroleum Institute, June 1990.

3. ASME B96. I : Welded Aluminum-Alloy Storage Tanks, The American Society of Mechanical Engineers, 1993.

4. AWWA 0100: AWWA Standard for Welded Steel Tanks for Water Storage, American Water Works Association, 1984.

13. AISC Manual of Steel Construction, 9th edition, American Institute of Steel Construction, Inc., 1989.

14. Section 1, Bulletin 30: Specifications for Aluminum Structures, 5th edition, The Aluminum Association, December 1986.

15.1994 Uniform Building Code, International Congress of Building Officials, May 1994.

Suggestions for further reading For more information on designing storage tanks, read “Part I: Development of design loads” of the authors’ original paper, “Bulk storage tank design considerations,” in Proceedings of the 1996 Powder & BulkSolids Confer- ence, May 1996, Rosemont, Ill.

See additional articles listed under “Storage” in Powder and Bulk Engineering’s comprehensive “Index to articles,” December 1995, page 89.

Gary A. Chubb, PE, is chief design engineerfor Peabody TecTank, PO Box 996, Parsons, KS 67357; 316/421- 0200. This article is adapted from Part II of a paper he presented at the 1996 Powder & Bulk Solids Conference in Rosemont, Ill. Keith E. McGuire, PE, is Peabody se- nior design engineer andpresented Part I of the papel:

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