For a vessel floating on the surface, the hydrostatic properties of relevance are the conditions of flotation and the stability of the vessel in relation to disturbances from the static flotation condition.

It is assumed that the water surface is calm and any movement of the craft is sufficiently slow for any dynamic effects to be discounted.

Accepting Archimedes principle that a vessel will displace its own weight of water, the initial requirement for flotation is that the intact hull is of sufficient volume to displace its own weight whilst having a reasonable freeboard (height of weather deck above the waterline) and that it floats upright. The volume of watertight hull above the waterline constitutes a Reserve of Buoyancy (ROB). This ROB becomes important when considering the safety of the vessel in damaged conditions when water floods part of the hitherto intact displacement volume below the waterline.

The stability of the vessel concerns the outcome of perturbations from the static flotation condition. Sideways motion (sway), change of heading (yaw) and fore or aft motion (surge) do not change the static conditions of the hull and can be considered as neutral. Whereas roll motion (heel), pitching and vertical motion (heave) result in changes in the distribution of buoyant volume and hence variation in the static equilibrium condition. The question to be answered is whether after a disturbance the vessel returns to its initial equilibrium state.

If a vessel heaves upwards then the displacement volume reduces and the excess of unchanged weight over reduced buoyancy provides a force in a direction restoring the vessel to its original position.

INTRODUCTION

8.1 Historically, it was not until the advent of the nuclear submarine with its capability for sustained high speed that the focus of attention of the submarine designer moved from the provision of adequate means for control of motion in the vertical plane – particularly at periscope depth – to achievement also of an appropriate balance between manoeuvrability and dynamic stability, with the emphasis again on motion in the vertical plane. Concentration of attention on depth changing and keeping rather than on course changing and keeping was natural because of the inherent risks in the high speed submarine – even with better diving depth capability through increased pressure hull strength – of accidentally exceeding the allowed maximum depth, a hazard regarded as potentially more dangerous than inadvertent surfacing.

As can be appreciated, there are similarities between the submarine manoeuvring submerged and the airship – though there are also significant differences – and in fact early theoretical and experimental investigations into submarine dynamic stability and control, in the 1940s and 1950s, initially drew on corresponding investigations for airships in the 1920s and 1930s. Subsequent research, specific to submarines, has provided a body of knowledge of which it behoves the submarine designer to have a basic understanding, even though dynamics and control are not primary considerations in determining the size and shape of the submarine at the concept stage.

It might seem from the latter observation that, once size and form have been determined by the dominant considerations, provision of the means for achieving the desired control characteristics could be left to specialist hydrodynamicists and control engineers for development subsequent to the concept stage.

INTRODUCTION

5.1 The main attribute of a submarine is its ability to dive beneath the surface and to go to reasonable operating depths. For a manned submersible there is a requirement for the enclosed volume to be maintained at atmospheric pressure. This need applies not only for the personnel but also for much of the equipment which has been designed to operate in atmospheric conditions. It is desirable to keep the enclosed volume as small as possible so as to limit the weight of structure that is required to withstand the differential pressure between sea pressure at depth and atmosphere. In small unmanned submersibles and ROVs it will usually be possible to minimise the amount of volume that needs to be contained by the pressure carrying structure, but for most large seagoing manned vessels there are inescapable requirements for a considerable amount of volume to be contained within the structural envelope. It may be that the design of a submarine as a whole leads to a decision to include other volumes within the pressure hull although they are not necessarily required to be at atmospheric pressure; for example, some of the main ballast tankage may be included within the pressure hull. It can also be convenient to locate some fuel tanks within the pressure hull so that they can be operated in atmospheric conditions. Variable ballast tanks will usually be located within the pressure hull, even though in some instances they are subject to sea pressure. It is important that where such volumes are inside the pressure hull, steps are taken to enable them to be isolated from sea pressure.

Introduction

Bubbles form in a flowing liquid in areas where the local pressure is close to the vapour pressure level. They form and collapse in a short time, measured in microseconds, and their life history gives rise to local transiently high pressures with flow instability. In pumps this results in noise, vibration and surface damage which can give rise to very considerable material loss.

The inception and collapse mechanisms are discussed briefly in this chapter, as are the conventional empirical rules used to ensure satisfactory pump behaviour. The chapter concludes with a discussion of the design rules to be followed in producing a good pump, and with a treatment of the techniques used to predict cavitation performance.

Bubble inception and collapse

In theory, cavities will form when the local liquid pressure level is equal to the vapour pressure under the local conditions. In practice bubbles form at higher pressure levels, due in part to the presence of very small bubbles or particles of detritus which act as triggers. A very exhaustive treatment of the process will be found in the monograph by Knapp et al. (1970), so a very brief summary will be given here.

Figure 2.1 is based on work done by Worster (1956) who used theoretical equations first published by Rayleigh (1917) to predict the life cycle of an existing small bubble as it grew and then collapsed.

This text is intended as an introduction to rotodynamic pump design. Any successful pump must satisfy the following objectives:

It must give specific pressure rise and flow rate within acceptable limits at an acceptable rotational speed, and take minimum power from its drive; it must give a stable characteristic over the operating range required and while meeting all performance criteria, the cavitation behaviour must be good. The pump must be as small as possible, the power absorbed must normally be non-overloading over the flow range and the noise and vibration must be within specified limits. The design must always be economical, give good quality assurance, and be easily maintained.

In addition to the objectives stated, special requirements also have an influence on design. For example, pumps handling solids must resist erosion and blockage of flow passages. In many fluid processes the pumps have to cope with the multiphase fluids and high gas content. Modern boiler feed pumps pose particular problems of shaft and drive design. This text introduces the reader to design approaches which can deal with these and other problems.

It has been assumed that the reader has a basic understanding of fluid mechanics, so the treatment commences with a statement of the fundamental Euler equation and its applications, and continues with a fairly comprehensive discussion of cavitation, its effects, and basic design data relevant to rotodynamic pumps. The text then describes the fundamental design principles and information available on centrifugal and axial/mixed flow machines.

Introduction

In the study of fluid mechanics, Newton's second law of motion enables the Eulerian equations of motion of a fluid to be applied to a study of the forces acting on a fluid particle at a particular point at a particular time. The solution of these equations with the true boundary conditions in a pump is a formidable task, because within a pump there are rotating and stationary blades that change in their orientation and cross-sectional geometry from hub to tip. There also are boundary layers on the annulus walls and the blade surfaces, wakes from the trailing edges of the blade, over-tip leakage flows etc, so the flow is unsteady, three-dimensional, and viscous.

A brief reference has already been made to empirical approaches in the chapter on axial and mixed flow machines principles. In this chapter the empirical approach to determining passage shapes will first be outlined, and then the more analytical techniques made possible by the computer will be outlined.

Stream-surfaces

Where pumps are of radial, of Francis type or completely mixed flow layout, the principles for centrifugal pumps already covered are often used, once the stream-surfaces are established. In the following sections the approach to the shape of stream surfaces is discussed, and then empirical solutions are outlined.

Stream-surface design

Solutions are described in some detail in texts such as that by Wisclicenus (1965) and in less detail by Turton (1984a).

Introduction

Chapters 1 to 7 have outlined the principles that underlie the successful design of the elements of the flow path in centrifugal, axial and mixed flow machines, and with the help of worked examples the logic to be followed has been explained. Chapter 8 has introduced the background to shaft design, seal and bearing system selection, and Chapter 9 has introduced the problems posed by difficult liquids, and the principles to be followed in choosing the correct materials and thickness of machine elements.

The first stage in the process of developing a pump is the establishing of the pump duty, which does not just mean the duty point but the fluid properties and other matters. Detail design is only complete when all the relevant standards and codes of practice are observed and satisfied. The discussion which follows covers these matters, and concludes with a brief discussion of test provisions and procedures.

Establishing the pump duty

An important factor in the process of producing the right pump design is the establishment of the pump duty. This demands a full interchange of information between the customer and the pump maker.

The pump designer requires to know, in addition to the rated flow rate, head and NPSHA, the environment in which the machine is to be operated, the probable range of flow rates and heads that are to be presented to the machine in the plant.

Introduction

Axial and mixed flow pumps will be treated as members of the same family, as they are high specific speed machines, even though the ‘Francis’ type of centrifugal machine has a mixed flow path. Both types exhibit the same characteristic behaviour, with a rise of specific energy towards shut valve which can be high in axials and in some mixed flow machines, and share a distinct tendency to unstable behaviour at part flow, and a power requirement which rises as flow reduces.

Energy is imparted to the fluid by blades rather than passages, and instability arises from flow breakdown over the blade profiles (giving rise to stall effects as described in Chapter 4). The interaction of fluid and pump components is complex, and performance is affected by blade profiles, surface finish, small variations in blade spacing and setting, and intake disturbances.

The principles underlying isolated blade profiles and blades in close proximity were described in Chapter 4, and simple but reasonably effective design techniques can be applied to the axial flow pump when the inlet flow is undisturbed. Quite efficient machines have been designed in the empirical way outlined in Section 7.3, but higher performance axial machines and all mixed flow machines require more sophisticated approaches, discussed in Section 7.4. Computer based methods were discussed in outline in Chapter 5.

Introduction

Typical flow paths for axial and mixed flow machines are shown in Figure 4.1. Axial flow pumps are usually fitted with a rotor only, so that there is very little pressure recovery after the impeller and even where outlet guide vanes are fitted their main function is to remove any outlet swirl from the flow. Mixed flow pumps may either be as shown in Figure 4.1(b) without outlet guide vanes, or as in many machines, for example in the bulb or bore hole pumps, guide vanes are fitted to improve the flow into the second stage of the assembly.

Unlike the centrifugal pump, the performance in axial machines in particular is a function of the action of blade profiles. Only in mixed flow pumps with many blades is the dominant fluid dynamic action that of the passages as in centrifugal machines.

The fundamental relations have been introduced in Chapter 1, and the application of the Euler equation was demonstrated. In this chapter data for isolated aerofoils is discussed, as it applies to axial machinery, and the concepts of radial equilibrium and stall are introduced. This material forms the basis of empirical design techniques, where it is assumed that all stream surfaces are cylindrical. This is only approximately true in axial machines, and in mixed flow machines it is necessary to establish a number of stream surfaces and then either use the axial data along each surface, or use more advanced analytic fluid dynamic solutions based on the surfaces. This chapter therefore outlines an approach to stream surface shape determination and to mixed flow empirical and analytic solutions.

Introduction

It is not possible to cover anything in this book but the most basic considerations that underly the choice of shaft bearing seal and drive, so that the reader is referred to the literature and to pump handbooks that give more information. In this chapter the very basic principles are introduced, and the conventional terms used are defined.

Shaft design is first introduced, it being commented that this must involve the whole rotating system. Rolling element and plain bearings are then discussed, the basic seal designs are introduced and some design rules for good service are outlined. The chapter concludes with references to the selection of drive arrangement.

Shaft design

The shaft in a pump must sustain torsional effects, bending forces due to both the mechanical parts and hydraulic loads, and axial loads due to weight in the vertical plane and to hydraulic loads.

The empirical approach to shaft design is well documented in such texts as Stepannof (1976), Karassik et al. (1976) and the engineering handbooks. If the weight of the impeller system is known, and the axial and radial hydraulic loads determined, the shaft sizes can be determined and checked. Consider Figure 8.1, showing the rotating assembly for a horizontal centrifugal pump. Simple statics allow the determination of the reactions R1 and R2 and the resulting moment applied to the bearing system can be calculated.

Introduction

When designing a pump a number of design variables need to be determined:

• impeller rotational speed

• impeller inlet or suction dimensions

• impeller outlet diameter

• impeller blade number

• impeller blade passage geometry, including inlet and outlet blade angles

• impeller position relative to the casing

• collector leading dimensions (volute throat area or diffuser geometry)

• pump construction and materials.

There are a number of approaches to design, chief among which are: small changes from existing designs to give a slight change in head or flow range; design using empirical information, tabular and graphical; and computer based approaches which are in some instances based on empirical data and more recently use finite element or finite difference approaches. The use of these techniques will be discussed later. The sections which now follow survey some of the empirical information available. Typical pump cross-sections of single-stage end suction, and double suction designs and of a multi-stage machine are shown in Figures 3.1, 3.2 and 3.3.

Choice of rotational speed

As will be clear from a reading of the later chapters on design, the choice of rotational speed is interlocked with other parameters, but there are empirical speed limits as given, for example, by the American Hydraulic Institute Standards (1983) reproduced in many handbooks. Clearly the rotational speed is limited to a range of synchronous speeds when using electric motor on a 50 or 60 HZ supply frequency. For large pumps, turbine or diesel drive is used, and the eventual rotational speed is a compromise between hydraulic design and driver considerations.