Russell Davison

I originally wrote this article, “Concepts of entrepreneurship” in March 2004.

Entrepreneur traits, creativity, innovation, business planning and growth management are five of the main concepts of entrepreneurship. Lists of characteristics common to entrepreneurs have been published by many authors but others suggest that previous experiences are more important. Entrepreneurial creativity requires a paradigm shift and there are many techniques available to help the entrepreneur to see things in a different perspective, to come up with new ideas. Innovation involves implementing newly created ideas and the process can be classified as invention, extension, duplication and synthesis. Strategic planning is used to assess the entrepreneur's position in external/internal environments, identify key success factors/competencies and to implement a strategy. Finally, the issue of growth management requires the entrepreneur to settle on what size of company he is happy with, how much direct control is afforded to him and how entrepreneurial spirit can be retained in a growing business.

ENTREPRENEUR TRAITS

Many authors have published lists of characteristics that they consider to be displayed by entrepreneurs. Cunningham and Lischeron (1991) have grouped these contributions into six schools of thought. They classify these as the; "Great Person", Psychological, Classical, Management, Leadership and Intrapreneurship schools of thought.

"Great person" - Born entrepreneurs, e.g. Fords, Rockefeller, Trump.
Psychological - Entrepreneurial personality, behaviour developed over time.
Classical - Entrepreneurial key factors are innovation and creativity.
Management - Entrepreneurs can be developed or trained in the classroom.
Leadership - Attract people to support a vision and transform it into reality.
Intrapreneurship - Encouraging people to work in semi-autonomous units.

However, much criticism is levelled at these theories because many of the characteristics are not unique to entrepreneurs and can be found in successful managers and executives. Liles (1974:43) proposes that "certain kinds of experiences and situational conditions - rather than personality or ego - are the major determinants of whether or not an individual becomes an entrepreneur" and Bailey (2003) questions whether entrepreneurs possess different characteristics or whether they are merely products of unique situational factors. This view is also supported by O'Neile (1989), who affirms that the entrepreneur is a "product of his historical and environmental circumstances." The choice to become an entrepreneur must be influenced by events that led to the decision, claims Brockhaus and Horwitz (1986). They suggest that previous experience has an effect. These previous experiences could be positive, such as role models and education, or they could be negative displacements. Refugees and migrants may choose entrepreneurship if gaining employment is difficult. Job dissatisfaction or job loss may be other stimuli to select entrepreneurship.

CREATIVITY

Entrepreneurship can be partly described as a combination of creativity followed by innovation, where creativity is the act of 'thinking' new things, coming up with ideas and innovation is 'doing' new things or implementing the newly created ideas. Creativity is also concerned with new ways of looking at opportunities and new approaches to solving problems. This may require the entrepreneur to shift paradigms and discard old assumptions and perspectives. Mukerjea (2003), in "Brain Symphony", describes sixteen techniques that can be used by entrepreneurs to stimulate creativity:

Visual Gym - Creating scenes through imagination, used by Nikola Tesla.
Torrence Tests - Reverse, substitute, modify, adapt, new uses, combine, eliminate, simplify.
Random' Riting - Paragraph creation from randomly selected words.
Cinquains - Noun, two adjectives, three verbs, four word statement, noun.
Matchmaking - Attribute matrices, linking, lists and morphological analysis.
Radiant Thinking - Word association to branches radiating from the centre.
Metaphorical-Analogical Thinking - Problem, analogy, attributes, emergent ideas.
Cut n' Paste - Collage of cut-out images with captions.
Abstract Designs - Creative interpretation of instructions for drawing objects.
Object Analogy - Use ordinary objects to draw analogies for problem solving.
Freewheeling - Combine randomly selected objects to produce new objects.
Mentamorphosis - Infusing oneself into the actual form of the central problem.
Ideavisuals - Picture codes and storyboarding, used by Wait Disney.
Kaleidoscoping - Mixing and matching synonyms of the key problem words.
SitSol Reversal - Reverse the situation and focus on the negatives.
Fishboning - Cause and effect diagram for clarifying ambiguities.

Another technique is to “surround yourself with people who are different from you. Always ask for help and another point of view - even when you may not think that you need it. You'll often be surprised that there is a better way to look at the original idea", says Gillian Franklin, according to Turner (2003). Once the entrepreneur has created, or discovered, new ideas then they are evaluated against each other as a candidate for innovation.

INNOVATION

Schumpeter (1934) identifies the entrepreneur's challenge as discovering and implementing new ideas. He asserts that innovation is a unique feature which separates entrepreneurs from managers. It's stated that this is achieved by (1) developing new products or services, (2) developing new methods of production, (3) identifying new markets, (4) discovering new sources of supply, and (5) developing new forms of organizations.

The innovation process can be categorized into four basic types, suggests Kuratko and Hodgetts (2004). These are; invention, extension, duplication and synthesis. Novel products or services are 'inventions', and the application of a current concept to a different application is an 'extension'. An improvement to an already existing concept is a 'duplication' and forming a new application from existing concepts is 'synthesis'.

BUSINESS PLANNING

Entrepreneurs are repeatedly monitoring windows of opportunity. These windows are continuously opening and closing and strategic planning is required to assess if the opportunity is worthwhile for the entrepreneur and how it should be successfully exploited. Whilst strategic planning is essential to ensure successful operation, it is a particularly useful tool when the entrepreneur's business is growing, it serves a niche market or business performance is improving. There are many schools of strategic management thought available to the entrepreneur and Mintzberg (1990:112) illustrates the Design School Model.

The Design School Model can be described as having eleven components:

External appraisal - An examination of the external elements influences the entrepreneur's strategy options. This involves investigating customers, competitors, market and the environment. Where the environment is political, economic, society, technology and ecology considerations.

Threats and opportunities in the environment - The external appraisal reveals the opportunities that the entrepreneur can exploit and the threats he faces. Opportunities are regarded as positive trends and threats are negative trends.

Key success factors - Key success factors are competitive assets or competences that the entrepreneur needs to compete successfully in his chosen industry. An absence of strategic necessities is a weakness and possession of strategic strengths will give advantage to the entrepreneur.

Internal appraisal - An examination of the skills of the entrepreneur's employees, resources, innovations and financial position discloses how the business is constrained by it's capabilities and resources.

Strengths and weaknesses of the organization - Any activities that the entrepreneur does well are identified as strengths from the internal appraisal. Any lack of resources or activities that the entrepreneur does not do well are identified as weaknesses.

Distinctive competencies or assets - Distinctive competencies are the activities that the entrepreneur does exceptionally well.

Social responsibility - Social responsibility is the entrepreneur's obligation, beyond that required by the law and economics, to pursue long-term goals that are good for society.

Managerial values - This describes how the entrepreneur's managers establish, promote and practice the business values. The building of team spirit, influencing marketing efforts, shaping of employee behaviour and guidance for manager's decisions and actions are examples of the main purposes of managerial values.

Creation of strategy - Strategic alternative strategies need to be developed by the entrepreneur for evaluation. These strategies should take advantage of environmental opportunities and exploit the company's strengths.

Evaluation and choice of strategy - Some of the criteria used for selection of a strategy from alternatives are scenario consideration, sustainable competitive advantage pursuit, organizational vision and objectives consistency, feasibility and the relationship to the other strategies of the entrepreneur.

Implementation of strategy - For the entrepreneur to succeed, the chosen strategy must be implemented and this involves converting strategic alternatives into an operating plan.

The final stage of business planning is to actually implement the strategic plan and four approaches are suggested. Strengths and weaknesses can be identified on a year-to-year basis in an opportunity management approach. Activities can be carried out sequentially in a milestone planning approach with achievable goals along the way. Expert theory can be used in a strategic model approach which states how the plan should be prepared and executed. Significant variables, venture phase and the entrepreneur's growth preferences can be applied to a contingency model.

MANAGING GROWTH

The entrepreneur's management of the growth of the business raises many important issues. Amongst these are; activity level, retaining entrepreneurial spirit, delegation, and ownership. The ability, need and opportunity will determine the business growth.

First and foremost, the entrepreneur has to make the big decision as to what level of activity he wants for the business. Churchill and Lewis (1992) categorize the entrepreneur's business growth into the five stages of existence, survival, success, takeoff, and maturity. So, the entrepreneur must want to move from each stage to the next, otherwise he must remain at a stage where he is comfortable with the level of activity.

The high entrepreneurial spirit that exists in a small business can erode as the business grows in size, unless a conscious effort is made to avoid the adverse effects of bureaucratization. This can be achieved by creating entrepreneurial momentum that exists within the business and which, in itself, becomes a driving force of entrepreneurship. The business cannot grow unless the entrepreneur is able to effectively delegate. This can create a problem for many entrepreneurs who might fear that they are losing control of the business by giving greater responsibility to their staff. The act of delegation is made easier if the entrepreneur is able to create work procedures for standard activities like accounting, dispatch, etc.

The final main issue of growth management is that of ownership. It is not possible for most entrepreneurs to achieve growth with their own limited financial resources and the surplus cash flow of the business, especially in the early years. Working capital requirements increase dramatically with increasing growth rates because of the need to outlay capital for expenditure before revenue is received. For this reason, the entrepreneur must raise capital through debt or dilution of ownership. Many entrepreneurs fear that shareholders or venture capitalists will influence the management of their venture or, in the extreme, remove them from the business.

References

Bailey, J. 2003, 'The right stuff', Business Review Weekly, 4-10 September, pp. 32-34, 36.

Brockhaus, R.H. & Horwitz, P.S. 1986, 'The psychology of the entrepreneurs' in Sexton, D.L. & Smilor, R.W. (Eds), The Art and Science of Entrepreneurship, Cambridge, MA: Ballinger Publishing pp. 25-48.

Churchhill, N.C. & Lewis, V.L. 1992, 'The five stages of business growth' in Sahlman, W.A. & Stevenson, H.H. (Eds), The Entrepreneurial Venture: Readings, Boston, MA: Harvard Business School Publications pp. 263-275.

Cunningham, J.B. & Lischeron, J. 1991, 'Defining entrepreneurship', Journal of Small Business Management, Vol. 29, January, pp. 45-59. Kuratko, D.F. & Hodgetts, R.M. 2004, Entrepreneurship: Theory, process, and Practice, 6th edn. Ohio: South Western.

Liles, P. R. 1974, 'Who are the entrepreneurs?', MSU Business Topics, Vol. 22, No. 1 pp. 43-55.

Mintzberg, H. 1990, 'Strategy Formation - Schools of Thought', in J.W. Frederickson (ed) Perspectives on Strategic Management, Harper Business, New York.

Mukerjea, D. 2093, Brain Symphony, Horizon Books pte Ltd, Singapore.

O'Neile, M.J. 1989, 'The entrepreneur in economic thought', Armidale, NSW: Department of Accounting and Financial Management University of New England.

Schumpeter, J.A. 1934, The Theory of Economic Development, Harvard University Press, Cambridge, Mass., p.56.

Turner, R. 27hey just did it', Boss Magazine, April, pp. 55-56, 59-60, 64-70.

I was asked to carry out this 16” angle control valve seismic analysis for one of the world’s leading engineering companies.

1.0 Summary of findings

2.0 Introduction

3.0 Identification of susceptible features to be analysed

3.1 Body studs

3.2 Area behind outlet flange

4.0 Identification of operational loads

4.1 Pressure of the line fluid

4.2 Tightening of the body stud nuts

4.3 Mass of the valve, actuator and bonnet

5.0 Calculations

5.1.0 Stresses imposed on the body studs by the action of operating loads and seismic acceleration, along the horizontal axis
5.1.1 Stresses caused by the hydrostatic end thrust or gasket seating
5.1.2 Bending stress due to the actuator
5.1.3 Bending stress due to the bonnet
5.1.4 Bending stress due to the body
5.1.5 Direct stress due to the actuator weight
5.1.6 Direct stress due to the bonnet weight
5.1.7 Direct stress due to the body weight
5.1.8 Direct stress due to the actuator thrust
5.1.9 Summation of stresses

5.2.0 Stresses imposed on the body studs by the action of operating loads and seismic acceleration, along the vertical axis
5.2.1 Stresses caused by the hydrostatic end thrust or gasket seating
5.2.2 Direct stress caused by the body under the influence of acceleration due to gravity plus seismic effects
5.2.3 Direct stress caused by the actuator under the influence of acceleration due to gravity plus seismic effects
5.2.4 Direct stress caused by the bonnet under the influence of acceleration due to gravity plus seismic effects
5.2.5 Direct stress due to the actuator thrust
5.2.6 Summation of stresses

5.3.0 Stresses imposed on the area behind the outlet flange by the action of operating loads and seismic acceleration along the horizontal axis
5.3.1 Longitudinal stress caused by the internal pressure
5.3.2 Bending stress due to the actuator
5.3.3 Bending stress due to the bonnet
5.3.4 Bending stress due to the body
5.3.5 Direct stress due to the actuator weight
5.3.6 Direct stress due to the bonnet weight
5.3.7 Direct stress due to the body weight
5.3.8 Direct stress due to the actuator thrust
5.3.9 Summation of stresses

5.4.0 Stresses imposed on the area behind the outlet flange by the action of operatingloads and seismic acceleration along the vertical axis
5.4.1 Longitudinal stress caused by the Internal pressure
5.4.2 Direct stress caused by the body under the influence of gravity and seismic effects
5.4.3 Direct stress caused by the bonnet under the influence of gravity and seismic effects
5.4.4 Direct stress caused by the actuator under the influence of gravity and seismic effects
5.4.5 Direct stress due to the actuator thrust
5.4.6 Summation of stresses

6.0 Results

1.0 Summary of findings

The 16" angle control valve will perform it's intended duty during an earthquake having a seismic acceleration of 0.81g in the horizontal direction, and during an earthquake having a seismic acceleration of 0.81g in the vertical direction. There will be no rupture of the valve body nor failure of the studs at the valve body/ bonnet joint.

2.0 Introduction

I consider the performance of a 16" angle control valve subjected to both horizontal and vertical seismic forces together with normal operating loads. Two regions are identified as being susceptible to damage during seismic events.

The maximum tensile stress is calculated for the two susceptible areas for normal operating loads plus seismic loading along the horizontal axis and for normal operating loads plus seismic loading along the vertical axis.

Component stresses are determined by accepted axial loading stress theory, bending stress theory, cylinder stress theory and bolt load theory.

These stresses are summated and compared with the yield stress of the material being considered to show that the total tensile stress is not in excess of the yield stress for non-pressurised components and not in excess of 90% of the yield stress for pressurised components.

The following assumptions are made in this analysis :

(1) The design seismic acceleration of 0.81g acts in the horizontal direction or vertical direction separately and not simultaneously.

(2) The pipe work and equipment adjacent to the valve exert no force or moment to it.

(3) The component stresses are all tensile and not compressive.

(4) The changes of section near to the region under consideration are such that negligable concentration of stress occurs.

3.0 Identification of susceptible features to be analysed

The two areas that we consider to be susceptible to seismic stresses are the body studs and the area behind the outlet flange.

3.1 Body studs


The body studs are manufactured in chromium-molybdenum steel ASTM A320 Grade L7. This material has a minimum yield stress of 105,000 psi and a minimum tensile strength of 125,000 psi There are 24 studs on a PCD (pitch circle diameter) of 30-1/2"

3.2 Area behind the outlet flange


The valve body is manufactured in carbon steel ASTM A350 LF2. This material has a minimum yield stress of 36,000 psi and a minimum tensile strength of 70,000 psi.

4.0 Identification of operational loads

The valve would be subject to stresses caused by the pressure of the line fluid, tightening of the body stud nuts and the mass of the valve body, bonnet and actuator.

4.1 Pressure of the line fluid

The maximum line pressure of the fluid is 1230 psig and this will act on all 'wetted parts'.

4.2 Tightening of the body stud nuts

A torque is applied to the body stud nuts which is sufficient to provide the body/bonnet seal by loading the body studs.

4.3 Mass of the valve body, valve bonnet and actuator

The mass of the valve body is 11,000 pounds and under the influence of acceleration due to gravity and earthquakes, this would produce a force and moment.

The mass of the actuator is 1,330 pounds and this also would produce a force and moment due to the above.

The mass of the valve bonnet is 1,200 pounds and under the influence of acceleration due to gravity and earthquakes, this would produce a force and moment.

5.0 Calculations

Firstly, we consider the stresses imposed on the body studs by the action of operating loads and seismic acceleration along the horizontal and vertical axis.

Secondly, we consider the stresses imposed on the area behind the outlet flange by the action of the above loads.

5.1 Stresses imposed on the body studs by the action of operating loads and seismic acceleration along the horizontal axis

The body studs are subject to stresses caused by the hydrostatic end thrust or gasket seating, bending due to the actuator, bending due to the bonnet, bending due to the body, force due to the actuator, force due to the bonnet, force due to the body and force due to the actuator thrust.

5.1.1 Stress causedby the hydrostatic end thrust or gasket seating

Using the method as described in Appendix 2 of the 'ASME Boiler and Pressure vessel code - Section VIII, Division I', calculations shall be made for each of the two design conditions of operating and gasket seating and the more severe will control.

The minimum required bolt load for operating conditions,

Wm1 = 0.785G^2 x p + (2b x 3.14 GmP)

where,

G = 27.050
P = 1,150 psig
b = 0.3186
m = 3.75

Wm1 = 0.785 x 27.050^2 x 1150 + (2 x 0.3186 x 3.14 x 27.050 x 3.75 x 1150)

Wm1 = 893,945 lbf

The minimum required bolt load for gasket seating,

Wm2 = 3.14 bGy

Where,

b = 0.3186 in
G = 27.050 in
Y = 7,600 psi

Wm2 = 3.14 x 0.3186 x 27.050 x 7,600

Wm2 = 205,663 lbf

As the larger of the two loads is due to the operating conditions, we consider the stress imposed on the studs due to hydrostatic end thrust,

S1 = Wm1 / A

Where,

Wm1 = 893,945 lbf

A = bolt area, 24 x 1.78 = 42.72 in^2

S1 = 893,945 / 42.72

S1 = 20,926 Ibf/in^2

5.1.2 Bending stress due to the actuator

The bending stress due to the actuator under the influence of seismic acceleration is given by,

S2 = (M / I) / Y

Where,

M = bending moment at stud, lbf in
= mass of actuator x seismic acceleration x distance
= (1330 / 32.2) x (0.81 x 32.2) x 59.25
= 63,830 lbf in

Y = distance from neutral axis to outermost fibre, in
= 16.063 in

I = second moment of area of twenty four studs, in^4

=(1 + AL^2)
= 24 x 0.2302 + 4 (1.78 x (1.9905^2 +5.8359^2 + 9.2836^2 + 12.0986^2 + 12.0986^2 + 14.08916^2 + 15.1195^2 ))
= 4,973 in^4

S2 = (63,820 / 4,973) x 16.063

S2 = 206 lbf/in^2

5.1.3 Bending stress due to the bonnet

The bending stress due to the bonnet under the influence of seismic acceleration is given by,

S3 = (M / I) Y

Where,

M = bending moment of stud, ibf in
= mass of bonnet x seismic acceleration x distance
= (1,200 / 32.2) x (0.81 x 32.2) x 4
= 3,888 lbf in

Y =16.063 in
I = 4,973 in^4

S3 = (3,888 / 4,973) x 16.063

S3 = 13 lbf/in^2

5.1.4 Bending stress due to the body

The bending stress dueto the body under the influence of seismic acceleration is given by,

S4 = (M / I) Y

Where,

M = bending moment at stud, Ibf in
= mass of body x seismicacceleration x distance
= (11,000 x 32.2) x (0.81 x 32.2) x 17.75

= 158,153 lbf in

Y = 16.063 in
I = 4,973 in^4

S4 = (158,153 / 4,973) x 16.063

S4 = 511 Ibf/in^2

5.1.5 Direct stress due to actuator weight

The direct stress imposed on the studs by the weight of the actuator is given by,

S5 = W2 / A

Where,

W2 = Weight of actuator, 1,330 lbf
A = bolt area, 42.72 in^2

S5 = 1330 / 42.72
S5 = 31 Ibf/in^2

5.1.6 Direct stress due to bonnet weight

The direct stress imposed on the studs by the weight of the bonnet is given by,

S6 = W3 / A

Where,

W3 = weight of bonnet, 1,200 lbf
A = bolt area, 42.72 in^2

S6 = 1200 / 42.72
S6 = 28 lbf / in2

5.1.7 Direct stress due to the body weight

The direct stress imposed on the studs by the weight of the body is given by,

S7 = W1 / A

Where,

W1 = weight of body, 11,000 lbf
A = bolt area, 42.72 in^2

S7 = 11,000 / 42.72
S7 = 258 Ibf/in^2

5.1.8 Direct stress due to actuator thrust

The direct stress on the studs caused by the thrust of the actuator when the valve is in the closed position and the valve is not filled with line fluid is given by,

S8 = F / A

Where,

F = maximum thrust of actuator, 15,600 lbf
A = bolt area 42.72 in^2

S8 = 15,600 / 4,272
S8 = 365 Ibf/in^2

5.1.9 Summation of stresses

Addition of the eight component stresses yields a total stress of :

S = S1 + S2 + S3 + S4 + S5 + S6 + S7 + S8
= 20,926 + 206 +13 + 511 + 31 + 28 + 258 + 365
= 22,338 lbf/in^2
. .
The value of the total stress is less than 22% of the yield-stress for this non-pressurised component.

5.2 Stresses imposed on the body studs by the action of operating loads and seismic acceleration along the vertical axis

The body studs are subject to stresses caused by the hydrostatic end thrust or gasket seating, force due to the actuator, bonnet, body and actuator thrust.

5.2.1 Stress caused by the hydrostatic end thrust or gasket seating

The direct stress caused by the hydrostatic and thrust or gasket seating calculation is as detailed in 5.1.1 yielding a stress of 20,926 lbf/in2.

5.2.2 Direct stress caused by the bodyunder the influence of acceleration due to gravity plus seismic effects

The direct stress is given by,

S9 = F / A

Where,

F = effective force due to body, Ibf

=11,000 + (11,000 / 32.2) x (0.81 x 32.2)

=19,910 lbf

A = bolt area, 42.72 in^2

S9 = 19,910 / 42.72
S9 = 466 Ibf/in^2

5.2.3 Direct stress caused by the actuator under the influence of acceleration due to gravity and seismic effects

The direct stress is given by,

S10 = F / A

Where,

F = effective force due to actuator, lbf

= 1,330 + (1,330 / 32.2) x (0.81 x 32.2)

= 2,407 lbf

A = bolt area, 42.72 in^2

S10 = 2,407 / 42.72

S10 = 56 lbf/in^2

5.2.4 Direct stress caused by the bonnet under the influence of acceleration due to gravity and seismic effects

The direct stress is given by,

S11 = F / A

Where,

F = effective force due to bonnet, lbf

=1,200 + (1,200 / 32.2) x (0.81 x 32.2)

= 2,172 lbf

A = bolt area, 42.72 in^2

S11 = 2,172 / 42.72

S11 = 51 lbf/in^2

5.2.5 Direct stress due to the actuator thrust

The direct stress on the studs, caused by the thrust of the actuator when the valve is in the closed position and the valve is not filled 2 with line fluid, is as detailed in 5.1.8 yielding a stress of 365 lbf/in^2.

5.2.6 Summation of stresses

Addition of the five component stresses yields a total stress of :

S = S1 + S9 + S10 + S11 + S8
= 20,926 + 466 + 2,407 + 51 + 365
= 24,215 lbf/in^2

The value of the total stress is less than 23% of the yield stress for this non-pressurised component.

5.3 Stresses imposed on the area behind the outlet flange by the action of operating loads and seismic acceleration along the horizontal axis

The area behind the outlet flange is subject to stresses caused by the internal pressure, moment of the actuator, moment of the bonnet, moment of the body, force due to the actuator, force due to the bonnet, force due to the body and force due to the actuator thrust.

5.3.1 Longitudinal stress caused by internal pressure

The longitudinal stress caused by the internal pressure is given by,

S12 = P (AI / AM)

Where,

P = Maximum working pressure, 1,150 psig
AI = inside areaof cross section in^2

= Pi x (14.75^2 / 4)

= 170.9 in^2

Am = metal area of cross-section in^2

= ( ((Pi x 19.25^2) / 4)) - ((Pi x 14.75^2) / 4)) )

= 120.1 in^2

S12 = 1,150 x (170.9 / 120.1)

S12 = 1,636 lbf/in^2

5.3.2 Bending stress due to the actuator

The bending stress due to the actuator, under the influence of seismic acceleration is given by,

S13 = (M / Y) I

Where,

M = bending moment at cross-section, Ibf in
= mass of actuator x seismic acceleration x distance
= (1,330 / 32.2) x (0.81 x 32.2) x 91
= 98,034 lbf in

Y= distance from neutral axis to outermost fibre, in
= 9.625 in

I= second moment of area at cross section, in^4

= Pi (19.25^4 - 14.75^4) / 64

= 4,417 in^4

S13 = (98,034 / 4,417) x 9.625

S13 = 214 lbf in^2

5.3.3 Bending stress due to the bonnet

The bending stress due to the bonnet under the influence of seismic acceleration is given by,

S14 = (M / I) Y

Where,

M = bending moment at cross -section, Ibf in
= mass of bonnet x seismic acceleration x distance
= (1,200 / 32.2) x (0.81 x 32.2) x 35.75

= 34,749 Ibf in

Y = 9.625 in
I = 4,417 in^4

S14 = (34,749 / 4,417) x 9.625

S14 = 76 lbf/in^2

5.3.4 Bending stress due to the body

The bending stress due to the body under the influence of seismic acceleration is given by,

S15 = (M / I) Y

Where,

M = bending moment at cross section, Ibf in
= mass of body x seismic acceleration x distance
= (11,000 x 32.2) x (0.81 x 32.2) x 20.375
= 181,541 lbf in

Y = 9.625 in
I = 4417 in^4

S15 = (181,541 / 4,417) lbf in x 9.625

S15 = 396 lbf/in^2

5.3.5 Direct stress due to the actuator weight

The direct stress due to the actuator weight is given by,

S16 = W2 / A

Where,

W2 = weight of actuator, 1,330 lbf

A = area of cross-section, 120 in^2

S16 = 1,330 / 120

S16= 11 lbf/in^2

5.3.6 Direct stress due to bonnet weight

The direct stress due to bonnet weight is given by,

S17 = W3 / A

Where,

W3 = weight of bonnet, 1,200 lbf
A = 120 in2

S17 = 1,200 / 120
S17 = 10 lbf/in^2

5.3.7 Direct stress dueto body weight

The direct stress due to the body weight is given by,

S18 = W1 / A

Where,

W = weight of body, 11,000 lbf
A = 120 in^2

S18 = 11,000 / 120
S18 = 92 lbf/in^2

5.3.8 Direct stress due to actuator thrust

The direct stress on the cross-section caused by the thrust of the actuator when the valve is in the closed position and the valve is not filled with line fluid is given by,

S19 = F / A

Where,

F = maximum thrust of actuator, 15,600 lbf
A = 120 in^2

S19 = 15,600 / 120
S19 = 130 lbf/in^2

5.3.9 Summation of stresses

Addition of the eight component stresses yields a total stress of,

S = S12 + S13 + S14 + S15 + S16 + S17 + S18 + S19
= 1,636 + 214 + 76 + 396 + 11 + 10 + 92 + 130
= 2,565 Ibf/in^2

The value of the total stress is less than 8% of the yield stress for this pressurised cross-section.

5.4 Stresses imposed on the area at he back of the outlet flange by the action of operating loads and seismic acceleration along the vertical axis

The area at the back of the outlet flange is subject to stresses caused by internal pressure, force due to the actuator, bonnet, body and actuator thrust.

5.4.1 Longitudinal stress caused by the internal pressure

The longitudinal stress caused by the internal pressure is the same as that calculated in 5.3.1 and is 1,636 Ibf/in^2.

5.4.2 Direct stress caused by the body, under the influence of acceleration due to gravity plus seismic effects

The direct stress is given by,

S20 = F / A

Where,

F = effective force due to the body, Ibf

= 11,000 + (11,000 / 32.2) x ( 0.81 x 32.2)

= 19,910 lbf

= 120 in^2

S20 = 19,910 / 120

S20 = 166 lbf/in^2

5.4.3 Direct stress caused by the bonnet under the influence of acceleration due to gravity plus seismic effects

The direct stress is given by,

S21 = F / A

Where,

F= effective force due to the bonnet, Ibf

= 1,200 + (1,200 / 32.2) x (0.81 x 32.2)

= 2,172 Ibf

A = 120 in^2

S21 = 2,172 / 120

S21 = 18 lbf/in^2

5.4.4 Direct stress caused by the actuator under the influence of acceleration due to gravity plus seismic effects

The direct stress is given by,

S22 =F / A

Where,

F = effective force due to the actuator, Ibf

= 1,330 + (1,330 / 32.2) x (0.81 x 32.2)

= 2,407 Ibf

A2 = 120 in^2

S22 = 2,407 / 120

S22 = 20 Ibf/in^2

5.4.5 Direct stress due to actuator thrust

The direct stress on the cross-section caused by the thrust of the actuator when the valve is in the closed position and the valve is not filled with line fluid is the same as that calculated in section 5.3.8 and being 130 Ibf/in^2.

5.4.6 Summation of stresses

Addition of the five component stresses yields a total stress of :

S = S12 + S20 + S21 + S22 + S19
= 1,636 + 166 + 18 + 20 + 130
= 1,970 Ibf/in^2

The value of the total stress is less than 6% of the yield stress for this pressurised cross section.

6.0 Results

The calculated total tensile stress in the body studs by the action of operating loads and seismic acceleration along the horizontal axis is 22,338 lbf/in^2. Increasing this calculated value by 25% gives 27,923 lbf/in^2 which is less than 27% of the material yield stress ( 105,000 lbf/in^2).

Under the action of seismic acceleration in the vertical direction the calculated total tensile stress in the body studs is 24,215 lbf/in^2. Applying the 25% increase produces 30,269 lbf/in^2 which is less than 29% of the material yield stress (105,000 lbf/in^2).

Considering the region behind the outlet flange, our calculations produce a total tensile stress in the section of 2565 lbf/in^2, for the action of operating loads and seismic acceleration along the horizontal axis. Increasing this calculated value by 25% gives 3206 lbf/in^2 which is less than 9% of the yield stress of the material ( 36000 lbf/in^2).

The calculated total tensile stress in the region behind the outlet flange, subject to operating loads and seismic acceleration in the vertical direction is 1970 lbf/in^2. Increasing this calculated value by 25% gives 2463 lbf/in^2, which is less than 7% of the material yield stress (36000 lbf/in^2).

I was asked to create this control valve cryogenic test procedure for one of the world’s leading engineering companies.

1 Introduction

2 Location of tests in manufacturing programme

3 Test temperature of minus (-)196 Celsius

4 Test temperature above minus (-)196 Celsius

5 Equipment specifications

1 Introduction

The company has the facility to carry out cryogenic testing of control valves at temperatures as low as minus (-)196 Celsius.  Tests conducted at these temperatures include:

(a) Seat leakage test
(b) Gland leakage test
(c) Body/bonnet joint leakage test
(d) Mechanical operation test

The tests are carried out on the valve when the required test temperature has been achieved and maintained for one hour.  This allows for a uniform temperature distribution throughout the assembly.  The allowable deviation from the desired temperature is ±3 Celsius degrees over this period.

There are two modes of operation for the cryogenic test plant, depending upon the required test temperature.  Heat transfer is effected by immersing the test valve in a bath of liquid nitrogen at its saturation temperature for a test temperature of minus (-)196 Celsius.If the required test temperature lies above minus (-)196 Celsius then an evaporator is used and this creates heat transfer by passing a mixture of evaporated nitrogen and cold air around the test valve.

2 Location of tests in manufacturing programme

All tests are carried out on the assembled valve and actuator upon completion of the hydrostatic and seat leakage tests.  Prior to testing, the valve is fully disassembled and degreased, ready for cryogenic testing.  The valve internal and external surfaces of the valve are to be free of moisture, dirt and metal particles.  The valve is fully disassembled, to check for any permanent distortion of components, upon completion of the test.

3 Test temperature of minus (-)196 Celsius

Blank flanges, with ¼” NPT connections, are bolted to the valve body and compress a PTFE gasket to provide an efficient seal.  Alternatively, for valves with butt welded ends, blind hubs are used.

The 1/4" metal tubing from the helium cylinders is attached to the inlet side of the valve. The tubing has a series of coils to increase the transfer of heat from the incoming helium, thereby reducing the temperature difference between the helium gas and the valve.  At the outlet of the valve, 1/4" metal tubing is connected to a bubble meter and flow meter.  Calibrated thermocouples are attached to the valve in five places :

(a) inlet blank flange (side)
(b) inlet blank flange (top)
(c) body/bonnet joint (body flange)
(d) body stud/nut
(e) actuator locking ring

The valve is carefully lowered into the insulated cooling tank, prior to the admission of liquid nitrogen.  A sufficient clearance must exist between the valve and the tank walls to allow for good circulation of the coolant.

Proof tests of the valve are carried out at ambient temperature to determine whether the test should continue and these tests are identical to those to be done at the test temperature.  The results of these tests should be better or equal to those required at the test temperature.

Upon satisfactory completion of the initial proving test, the valve is now ready to be cooled down to (-)196 Celsius.  Liquid nitrogen is poured into the insulated tank up to the body/bonnet joint.  Initially, rapid vaporisation of the nitrogen takes place, making it necessary to frequently pour liquid nitrogen into the vessel.  As the valve temperature is reduced to near the test temperature, only occasional 'topping up' is required to maintain the liquid nitrogen slightly above the body/bonnet joint.

When the test temperature has been achieved (± 3 Celsius degrees), a period of one hour must elapse to allow for the valve internals to reach the required temperature.  A purge of helium gas is required to be passed through the partially open valve to prevent the ingress of moisture throughout the cooling down operation.

The position of the valve is chosen so that the packing box, located in the valve bonnet, is clear of the cold boil-off gas.

(A) Seat leakage test - With the valve in the fully closed position, the helium gas pressure is raised to the seat test pressure of the valve.  The seat test pressure is equal to the maximum shut-offpressure in service.  This pressure may not exceed the cold pressure rating of the valve.  The valve trim leakage (if any) is recorded over a ten minute period and must not exceed a previously agreed value between the client and the company. Typical values for single seated globe valves, expressed as a percentage of the maximum flow coefficient (Cv), may be:

(a) 0.1000% metal piston rings
(b) 0.0060% nitrile '0' rings, balanced design
(c) 0.0050% p.t.f.e. lipseals
(d) 0.0020% metal/metal (standard)
(e) 0.0002% metal/metal (special lapped), solid design
(f) 0.0001% soft face
(g) bubble tight (special)

(B) Gland leakage test - With the body pressurised to the seat leakage test pressure, a 50% (by volume) solution of teepol/alcohol is brushed on to the top of the packing box.  Gland leakage, detected by the formation of bubbles, is not allowed. It may be necessary to slightly adjust the packing box flange nuts to achieve zero leakage because of component contraction during 'cool-down'.

(C) Body/bonnet joint leakage test - With the body pressurised to the seat leakage test pressure, careful examination of the body/bonnet joint area is undertaken.  The body/bonnet joint is submerged below the surface of the liquid nitrogen and any escape of helium gas (evidenced by bubbles) is not acceptable.

(D) Mechanical operation test - With no pressure in the valve body, i.e. just sufficient gas to prevent the ingress of moisture, the valve is stroked twenty times.  The valve should travel over its entire stroke length ina smooth action, with no jerks or jumpy action.

4 Test temperature of above (-)196 Celsius

The test valve is firmly supported on a platform above the evaporator.  Initial proving tests are carried out at ambient temperature to determine whether to proceed with the cryogenic test.

Upon satisfactory completion of the proving test, the evaporator is filled with liquid nitrogen.  A multi vane centrifugal fan draws air from the atmosphere and forces it through a heat exchanger.  This creates a reduction in the temperature of the air and evaporation of the liquid nitrogen.  The mixture of air and nitrogen then flows past the test valve in such a manner as to allow heat from the test valve to warm the mixture again, thereby reducing the temperature of the test valve.  The temperature of the test valve is regulated by the action of the centrifugal fan.

Various tests, as described in Section 3, are then carried out on the valve at the test temperature.

5 Equipment specifications

Fan
Type : 9” diameter Halifax multi-vane centrifugal fan
Outlet velocity : 4880 feet per minute
Speed : 2300 revolutions per minute
Static pressure : 1.75” water gauge

Heat exchanger
Tube diameter : 6” nominal
Tube length : 29 feet immersed
Evaporator Volume : 50 cubic feet

Insulated tank
Inside width : 1 foot
Inside length : 3 feet
Inside height : 2 feet

Overhead crane facility
Safe working load : 5 tonnes

Insulated enclosure
Insulation : 6" thick expanded polystyrene block
Housing : softwood
Inside width : 6 feet
Inside length : 6 feet
Inside height : 6 feet

Insulated pit
Insulation : 6" thick expanded polystyrene block
Sides : Concrete
Inside width : 6 feet
Inside length : 6 feet
Inside height : 6 feet

Electrical equipment
Five thermocouples
Digital voltmeter

Flow equipment
Alcohol Bubble meter
Positive displacement meter

I was asked to create this control valve mechanical test procedure for one of the world’s leading engineering companies.

1 Introduction
2 Location of tests in the manufacturing programme
3 Stem position error test
4 Deadband test
5 Hysteresis test
6 Hysteresis plus deadband test
7 Stroking time test
8 Operation instructions
9 Equipment specifications

1 Introduction

The company has the facility to provide a permanent record of an assembled valve’s performance, whilst undergoing mechanical operation tests. The results of these tests are recorded on A3 size paper using an XYT pen recorder.

2 Location of tests in the manufacturing programme

All tests are carried out on the assembled valve and actuator before hydrostatic and seat leakage testing. The packing box gland flange nuts are finger-tight and careful assembly of the valve ensures that the packing material is in an uncompressed state. A light lubricating oil is applied to the area of the stem that passes through the packing box.

3 Stem position error test

The purpose of the stem position error test is to verify that the desired and actual stem positions are within acceptable limits. The acceptable stem position error for the company’s range of control valves is ± 5% of the rated travel. The stem position should lie between the following upper and lower limits for an input signal of 3 to 15 psig:

INSTRUMENT	% OF RATED TRAVEL
AIR PRESSURE	LOWER	UPPER
(PSIG)	LIMIT	LIMIT
3		5.00
4	3.33	13.33
5	11.67	21.67
6	20.00	30.00
7	28.33	38.33
8	36.67	46.67
9	45.00	55.00
10	53.33	63.33
11	61.67	71.67
12	70.00	80.00
13	78.33	88.33
14	86.67	96.67
15	95.00

The equipment is connected to the valve and actuator as described in section 8.

4 Dead band test

Deadband is the range through which an input can be varied without initiating observable response. In a diaphragm actuated control valve, deadband is the amount that the instrument air signal can be changed without initiating valve stem movement.
The amount of deadband is determined by measuring the changeover pressure for a given stem position. The stem is taken up to a position of 25% of the rated travel. While stationary, the change in pressure which causes a change in stem movement is measured. The test is repeated at positions of 50% and 75%. This changeover pressure is known as the 'deadband' and it should not exceed the following values :

VALVE AND DIAPHRAGM ACTUATOR WITHOUT POSITIONER
SPRING RANGE	POINT	MAXIMUM CHANGEOVER
	OF TEST	PRESSURE
3 TO 15 PSIG	25%	0.20 PSI
	50%	0.25 PSI
	75%	0.35 PSI
6 TO 30 PSIG	25%	0.40 PSI
	50%	0.50 PSI
	75%	0.70 PSI

VALVE AND DIAPHRAGM ACTUATOR WITH POSITIONER
SPRING RANGE	POINT	MAXIMUM CHANGEOVER
	OF TEST	PRESSURE
3 TO 15 PSIG	25%	0.0056 PSI
	50%	0.0068 PSI
	75%	0.0096 PSI
6 TO 30 PSIG	25%	0.0112 PSI
	50%	0.0140 PSI
	75%	0.0192 PSI

The equipment is connected to the valve and actuator as described in section 8.

5 Hysteresis test

Hysteresis is a characteristic of a control valve that is the dependence of the stem position, for a given variation of the instrument signal, upon the history of previous variations and the direction of the varying instrument signal, i.e. increasing/decreasing. The amount of hysteresis is determined firstly by performing the deadband test, followed by stroking the control valve over its full travel and returning it to its starting point. The amount of hysteresis is calculated by deducting the deadband from the distance between the cyclic envelope at 25%, 50% and 75% travel. The hysteresis should not exceed the following :

POINT	MAXIMUM
OF TEST	HYSTERESIS
25%	0.40 PSI
50%	0.35 PSI
75%	0.25 PSI

The equipment is connected to the valve and actuator as described in section 8.

6 Hysteresis plus deadband test

Hysteresis plus deadband is the total dynamic friction present in a control valve and is the vertical or horizontal distance between the cyclic envelope obtained from the hysteresis test.  The acceptable hysteresis plus deadband for a control valve is ±5% of the rated travel.

The equipment is connected to the valve and actuator as described in section 8.0.











7 Stroking time test
The stroking time of a valve is the time taken for the valve to stroke over its entire travel. This may be from the fully open position to the fully closed position, or vice versa. The duration is measured from signal increase/ decrease to full travel.  As the stroking speed is dependant upon many factors, it is not practical to define acceptable limits. The influencing factors listed in order of priority are :

(a) Actuator size
(b) Actuator stroke
(c) Air supply
(d) Pressure
(e) Size of pipework connections
(f) Spring rate
(g) Air to open/close
(h) Type of positioner

The equipment is connected to the valve and actuator as described in section 8.

8 Operation of the test equipment

The plugs and sockets of the electrical equipment are individually numbered for ease of assembly and to eliminate the possibility of incorrect wiring. Plugs and sockets, having the same number, should be connected together - with careful consideration of the actuator fail position. There are two leads labelled 'R', and a further two labelled 'D'. When testing equipment having a reverse acting actuator, the leads labelled 'R' should be connected and when testing equipment having a direct acting actuator, the leads labelled 'D' should be connected.

The power supply for the pressure transducer is situated on the left of the cabinet and it is labelled 'PRESSURE TRANSDUCER'. The operating voltage of this power supply should be set at 10 VDC and this is achieved by careful adjustment of the coarse and fine potentiometers. In most cases, the pressure transducer power supply will already be set at exactly 10 VDC.

The power supply for the linear potentiometer is situated to the right of the pressure transducer power supply and it is labelled 'LINEAR POTENTIOMETER'. The operating voltage of this power supply should be set at 24 VDC and this is achieved by careful adjustment of the coarse and fine potentiometers. In most cases, the linear potentiometer power supply will already be set at exactly 24 volts VDC.

The scale setting for the instrument signal axis is situated on the left hand side of the pen recorder and it is labelled 'INSTRUMENT SIGNAL'. Three controls are required to be set and their positions are dependant upon the maximum instrument signal pressure used.

For a maximum instrument signal pressure of 15 psig, set the range knob to 2mV/cm, vernier to (approximately) 0.00 - 0.40, zero adjustment to (approximately) 4.90 - 5.10. The vernier and zero settings are approximate and may require fine adjustments to achieve full scale deflection.

For a maximum instrument signal pressure of 30 psig, set the range knob to 2 mV/cm, vernier to (approximately) 7.08, zero adjustment to (approximately) 4.19. The vernier and zero settings are approximate and may require fine adjustments to achieve full scale deflection.

The scale setting for the valve stroke axis is situated on the right hand side of the pen recorder and is labelled 'VALVE STROKE'. Three controls require setting and their positions are dependant upon the maximum valve travel.

For a maximum valve travel of 1.1/8", set the range knob to 0.1 V/cm, vernier to(approximately) 6.72, zero adjustment to (approximately) 5.20.
For a maximum valve travel of 1.1/2", set the range knob to 0.1 V/cm, vernier to (approximately) 7.63, zero adjustment to (approximately) 5.20.
For a maximum valve travel of 2.1/4", set the range knob to 0.2 V/cm, vernier to (approximately) 6.72, zero adjustment to (approximately) 5.20.
For a maximum valve travel of 3.1/2", set the range knob to 0.5 V/cm, vernier to (approximately) 1.76, zero adjustment to (approximately) 5.11.

All of the above settings for the vernier and zero adjustments are approximate only and may require fine adjustments to achieve full scale deflection.

9 Equipment specifications

Linearity of the linear potentiometer is better than 1%, as is the linearity of the pressure transducer.

EQUIPMENT	MODEL NUMBER	SERIAL NUMBER
Pen recorder - Farnell	RW101	F2058
Power supply (Potentiometer)	E30/1	005949
Power supply (Transducer)	E30/1	006189
Pressure transducer - Honeywell	136PC30G1
Linear potentiometer - Penny & Giles	LP26/200/6"/6K	108112B

This is a design proposal that I was asked to carry out for a Swedish world leading manufacturer of compressors, generators, construction and mining equipment, industrial tools and assembly systems. They required a magazine system to present end-pieces for the pre-assembly of pneumatic cylinders.

The second stage for the robot assembly of pneumatic cylinders involves the sub-assembly of end-pieces and half-pistons. End-pieces need to be handled by a magazine system because they are too large for conventional vibratory feeders. A magazine system is required at the pre-production facilities that is a scaled-down version of the future production system, within budget limitations. The cost of the system is split between the fixed cost for the transfer of parts to the robot and the variable cost of end-piece storage. The variable storage cost is proportional to the capacity of the magazine system. There is also an indirect labour cost for the filling and transport of magazines, in addition to the equipment material cost. The prototype can have the same transfer device as the production model, but with a smaller capacity.

MAGAZINE FILLING

The only economical method of magazine loading is to fill them at the point of final manufacture. This is because the time taken to insert a part into a magazine can approach the time taken to insert it into the part-built assembly. Nevertheless, end-pieces have to be transported from manufacture to assembly and magazines are the best way of doing this, whilst also giving protection to the surface finish.

MAGAZINE CAPACITY

The capacity of the magazine is as large as possible to achieve the minimum number of journeys from manufacturing to assembly during the shift. If demand for each cylinder diameter is equal then the magazine must contain in excess of sixty parts for a refill only once a shift. A single vertical stack magazine would be in excess of three metres high. It is therefore proposed that a number of units should be combined to form one magazine. Three magazines of twenty end-pieces seems reasonable.

PROPOSED SYSTEM

The system shown is one method of end-piece distribution. The illustration shows one magazine to store one style of end-piece. The production version for the Swedish manufacturing plant would have three magazines per end-piece, each behind the another.