i

LOW COST PELTON TURBINE DESIGN AND TESTING

Report Number: H/03/00074/REP

URN 03/1246

Contractor

Hydroplan UK Gilbert Gilkes & Gordon Ltd.

The work described in this report was carried out under contract as part of the DTI New and Renewable Energy Programme. The views and judgements expressed in this report are those of the contractor and do not necessarily reflect those of the DTI.

First published 2003 Gilbert Gilkes & Gordon Ltd. 2003

ii

iii

EXECUTIVE SUMMARY E1 Background to the Project

It is recognised that Pelton turbine technology may be used effectively over a relatively wide range head and flow conditions when compared to other turbine categories and is suitable for many medium and high head sites. The technology is proven and the overall concept is reasonably simple.

The purpose of this project has been to research, design and partially test two key components of a low cost Pelton turbine suitable for smaller UK sites and the export market. These components are the concentric casing surrounding the turbine and the turbine runner, specifically investigating bolted runner bucket technology.

From a previous assessment of the Gilkes range of Pelton turbines, two key facts have emerged from this work:- • The vertical shaft Pelton turbine configuration offers the most flexibility (and potential

least cost). • The runner represents 40% of the turbine cost.

This project has concentrated on both of these facts by investigating and developing:- (i) The Concentric Case Vertical Shaft Pelton (CCVSP). (ii) Bolted Runner Bucket (BRB) technology.

E2 Aims & Objectives of the Project The projects aim was to achieve the following specific targets:- • To research and investigate two specific areas of alternative design for Pelton turbines

and consider the hydro generating unit as a standardised product • To use computer aided design methods to analyse options in each area • To produce viable designs for prototype manufacture • To factory test the prototype design modifications • Ultimately, to engineer substantial cost savings into a flexible, standardised Low Cost

Pelton turbine which itself will enable savings to be made in the powerhouse construction

E3 Bolted Runner Bucket Technology The development of the BRB during the project has resulted in an innovation that has the potential to lead to substantial cost savings in runner manufacture. A UK patent has been applied for to protect the invention. This will be extended to cover each country where a competitor operates. A thorough testing program has shown the BRB will reduce manufacturing time and improve production control without compromising on operating performance.

iv

Cost / Benefits Assessment The new BRB design has the potential to lead to considerable savings in turbine manufacture while potentially leading to improvements in performance.

Manufacturing cost saving The new BRB design will dramatically reduce runner-manufacturing time. It was calculated that the first runner manufactured will save 55% of the cost of a conventional runner of the same capacity. This figure includes tooling that can be reused for subsequent machines of the same size. The estimated saving for subsequent units could be approximately 77% of the cost of conventional runner.

Manufacturing time A significant part of finishing a conventional runner is the bucket grinding and polishing process. BRB technology will dramatically reduce this time by allowing easy access to all the surfaces to be machined and the new process eliminates all hand finishing operations. The lead-time to manufacture a runner is reduced considerably. If the runner is entirely of a new design and solid modelling of the buckets and tooling are required, a reduction in time of 68% can be expected. If new tooling only is required, a reduction in lead-time of 79% can be achieved. A replacement runner can be made to an existing design with an 89% saving in lead-time. Hydraulic performance It is envisaged that hydraulic performance will improve due to greater control over the bucket surface geometry and finish. Fatigue considerations Historically, the problem with bolt-on buckets has been cyclic fatigue and corrosion exacerbating stress concentrations around the bolt leading to premature failure. The massive number of applied stress cycles, as high as high as 5x1010 over an operating lifetime, has been found to make fatigue failure the most likely failure mechanism. The FEM computer program has the capability to predict fatigue performance and reliable texts on the subject have been consulted, however the true test for the runner will be field trials.

v

Figure E1 - Assembled Test runner

E4 Concentric Case Vertical Shaft Pelton The new concentric case (CCVSP) has the potential for significant project benefits in three areas:- • Saving in powerhouse floor area required because of reduced manifold footprint. • Reduction in manifold construction costs because of greater simplicity. • Reduction in Headloss compared to standard manifolds An analysis was undertaken based on the layout of an existing 4 jet vertical shaft Pelton turbine installed by Gilkes in 1986. The conventional manifolds construction cost, performance and footprint area were then compared to that of a option 1 type manifold of the same capacity. CFD was used to model the existing and theoretical cases.

Cost comparison Prices were sought to manufacture the existing concentric case in the option 1 type design. The CCVSP was increased in capacity so that velocities in both manifolds were the same. Figure E2 below shows the existing manifold layout with the CCVSP (in blue) superimposed over it for comparison. Table E3 provides a comparison of performance.

vi

Figure E2 Footprint comparison

Jet Existing manifold - ∆h (m)

New concentric case - ∆h (m)

1st 2.0m 0.94m

2nd 2.25m 0.99m

3rd 3.5m 0.96m

4th 3.5m 1.01m

Table E3 Hydraulic performance

vii

Cost / Benefits The CCVSP design seems to have improved the turbine in three areas. • A reduction in powerhouse construction costs of 2 to 3% may be possible due to the

reduced footprint area required by the CCVSP. • Manifold costs may be reduced by 14% • A modest reduction in headloss may be possible and may improve further if flow baffles

were incorporated into the branch pipe transitions where the major flow separation occurred. Baffles have successfully used where vortex formation has been a problem.

E5 Conclusions The partnership between the leading UK turbine manufacturer and a leading UK small Hydro Consultancy has lead to the completion of a highly successful project . The BRB and CCVSP technologies are both novel and practical solutions to old problems. Further testing of both technologies will result in increased confidence which will in turn lead to a competitive market advantage for UK companies.

viii

1 INTRODUCTION 1.1 Background to the Project

It is recognised that Pelton turbine technology may be used effectively over a relatively wide range head and flow conditions when compared to other turbine categories and is suitable for many medium and high head sites. The technology is proven and the overall concept is reasonably simple. The purpose of this project has been to research, design and partially test two key components of a low cost Pelton turbine suitable for smaller UK sites and the export market. These components are the concentric casing surrounding the turbine and the turbine runner, specifically investigating bolted runner bucket technology.

From a previous assessment of the Gilkes range of Pelton turbines, two key facts have emerged from this work:- • The vertical shaft Pelton turbine configuration offers the most flexibility (and potential

least cost). • The runner represents 40% of the turbine cost.

This project has concentrated on both of these facts by investigating and developing:- • The Concentric Case Vertical Shaft Pelton (CCVSP). • Bolted Runner Bucket (BRB) technology.

1.2 Aims & Objectives of the Project The projects aim was to achieve the following specific targets:- • To research and investigate two specific areas of alternative design for Pelton turbines

and consider the hydro generating unit as a standardised product • To use computer aided design methods to analyse options in each area • To produce viable designs for prototype manufacture • To factory test the prototype design modifications • Ultimately, to engineer substantial cost savings into a flexible, standardised Low Cost

Pelton turbine which itself will enable savings to be made in the powerhouse construction

1.3 Modelling tools used in study

Computational Fluid Dynamics Overview

Computational Fluid Dynamics (CFD) is a computer-based tool for simulating the behaviour of systems involving fluid flow, heat transfer and other related and other related physical processes. It works by solving the equations of fluid flow (in a special form) over a region of interest, with specified (known) conditions on the boundary of that region. CFD was used extensively in the design of the CCVSP. The CFD program CFX 5.51 was used, which is well suited to modelling complex internal flow because of its advanced meshing features and solvers.

Finite Element Analysis

Finite Element Method (FEM) and allows stress, thermal and dynamic analyses to be performed. FEM produces design data and it produces this data quickly, easily and cheaply. It shows if a newly designed structure will work or not, and where the design needs to be improved. FEM was used extensively in the design of the buckets and runner for the BRB.

Solid modelling

Solid modelling was used extensively to develop the models for finite element analysis and to produce the machining data for the hard tooling required for wax pattern production.

2 BOLTED RUNNER BUCKET DESIGN & MODELLING

2.1 Gilkes Standard Turbine Range

Gilkes Water Turbine range is covered by four basic machine designs. • Pelton • Turgo Impulse • Francis • Kaplan For each type, there are a number of different specific speed variations enabling a large head and power output range to be achieved. Figure 2.1 illustrates how the four types fit into the head and power output envelope.

Low Cost Pelton Range

Figure 2.1 shows the range envisaged for Low Cost Pelton (LCP). It is quite conceivable that the BRB technology could be developed for a much higher power than shown in this figure. The CCVSP design is unlikely to be used on a machine of greater than 400kW. In fact the target range could be the 50 – 200kW range.

Basic prototype parameters

The Gilkes P316 model runner bucket with a PCD of 325mm and a bucket width of 121.4mm was selected for the low cost turbine modelling. This runner would produce 150kW under a head of 150m and a flow of 114 l/s. These parameters are not based on any particular site but were selected because they are typical values and are within the range indicated in figure 2.2 below.

Figure 2.1

Figure 2.2

2.2 Summary of existing BRB technology

Design Requirements Pelton runners are subject to a combination of stresses caused by centrifugal force and cyclic loads. The centrifugal force is induced by the by the fast rotating body and is related to the runner speed and mass. The cyclic load is induced each time the bucket meets the jet. On high-speed multi jet runners, the individual buckets can undergo as much as 5 x 1010 repetitive cycles over its lifetime. Therefore, runner susceptibility to fatigue failure must be a major consideration in the design process.

Other desirable runner features can be summarised as:-

• The absence of pre-induced stresses from welding or casting • Ease of final machining the buckets surfaces • High final surface finish to reduce hydraulic friction losses • High strength material with good wear and fatigue characteristics

Designs where interchangeable buckets are individually attached to the central runner disc are also desirable but is not widely used for a variety of reasons, but is mainly due to difficulties in avoiding premature fatigue failure. Several construction methods are currently employed in the manufacturing industry which provide an alternative to one piece castings.

Recent Developments

Interlocking structure pelton wheel

Fir Tree rooting or dove tailing is often used on steam turbines. However, steam turbines are exposed to very different forces compared to water turbines. In steam turbines, fir tree rooting allows movement and absorbs energy when resonance occurs. This movement in a water turbine may lead to fretting failure. An example of this concept has been developed by the German company EFG. The method is based on a forging technique to produce buckets that include a patented interlocking clamping system, which link together the buckets to form the runner. Individual buckets can then be individually removed for repair or maintenance. After forging, the buckets are individually CNC machined to improve the surface finish and complete the required tolerances.

Fig. 2.3 EFG Interlocking Runner Buckets

Weld deposition runners

The introduction of specialist robot controlled welding tools have made it possible to construct runner buckets from the accurate placement of weld deposits. Generally, the runner is based around a machined or forged disc, to which material is precisely added by the gas metal arc welding procedure. Part of the runner bucket is often partially machined from the central disc beforehand. All aspects of the deposition process are computer controlled. After the bucket shapes are completed, machining is still necessary to maintain surface finish and tolerances.

Fabricated runners

A combination of welding, forging and machining has been used successfully to manufacture Pelton runners. One method to construct runners is based on forged buckets integral with a tapering wedge. The wedge section is welded to other wedges to form the runner disc. The bucket end of the forging is only partially formed to allow ease of machining of the bucket profile. The remaining parts of each bucket are then hot forged from plate, which is welded to each partially complete bucket to complete the bucket shape. Final machining is required to provide a constant surface finish.

One piece machining of runner

The ultimate method of manufacturing a Pelton runner without the imperfections caused by welding or casting is to machine the runner from a solid forged disc. Firstly, the runner is modelled on a three dimensional CAD system. The completed model can be read by an advanced CNC milling machine which machines the disc into the required shape using

advanced tools to reach remote bucket areas. Final surface finishing is still required using hand techniques to achieve the required tolerance and surface finish.

Hooped Pelton runner

This runner overcomes the large bending stresses applied to overhanging buckets by introducing a supporting ring around the runner circumference. Two rings are attached to each bucket, which has been calculated to significantly reduce fatigue failures. The rings are located to fully support the runner while not significantly impeding the passage of fluid flow around the bucket.

Buckets Bolted into the Disc

This solution is where buckets are manufactured with an extension to their base, which then when placed together act like the “segments of a cake” building up the shape of the runner. There is also a centric grove milled into these segments, which then fits with a flat circular plate placed at either side, a singular bolt is then used to bolt the buckets into position. This was used in a vertical machine in La Rasse, in the Swiss Alps on a head of 479m and a flow of 180 l/s. Again, this is obviously feasible and can withstand relatively high heads, it is claimed that grinding was not required. However, the manufacturing tolerances must be very tight, requiring an expensive specialist machining capability.

2.3 BRB Hub Design

Maximum stress in the runner was set to about 20 – 30 MN/m2. For reference, Nechleba gives 30 MN/m2 normal stress, though runner materials are not specified. A value of 45 MN/m2

peak is recommended by Brekke. To use Brekke’s figure the surface defects must be no greater than 2mm by 1mm and subsurface defects no greater than 2mm by 2mm. The starting point was the need to attach a ring of buckets to the shaft. Previous work suggested that buckets could be attached to a runner using two side plates with bolts passing through wedge shaped extensions to the buckets, roots. Considering the physical aspects it was obvious that the roots were too narrow (about 0.031m) to carry reasonable sized bolts. It was decided to use wedge shaped roots, locked in the hub with tapered shoulders. The hub was modelled by the finite element method. Axisymmetric elements were selected to reduce the 3D problem to 2D. The ring of buckets were modelled as a solid disc because, at runaway, it is not expected that the joints between adjacent buckets will open. After many calculations, the optimum geometry appears to be as illustrated in figure 2.4.

.

Figure 2.4. Section showing selected hub and root geometry, stress due to runaway speed, 2776 rpm, 159 radians / sec The maximum principal stress in the model is 15.3 MN/m2 This occurs in the centre of the root and is caused by the centrifugal force trying to bend the shoulders back.

2.4 BRB Bucket design For production, it is intended to use lost wax casting techniques to produce the individual buckets. Reproducibility of the process is reported to be excellent and rework of the pattern after a test cast and dimensional check will reduce the deviation from drawing (Reference ASM Handbook Volume 15 page 248, 1998 [7] ). The prototype bucket was then dimensionally examined in detail at Gilkes’ factory by a co-ordinate measuring machine and checked against the drawing. Deviations were added to the pattern solid model to correct the geometry and a new, updated, pattern produced, this achieved a near net shape casting with the optimum hydraulic surface geometry and the minimum machining allowance.

2.5 BRB Shaft attachment

Several methods of connecting the hub to the shaft were investigated. Keying, shrinking and splining were rejected on cost or performance grounds. Clamping was investigated due to previous good experience with these devices. Eventually a technically acceptable design was evolved between Gilbert Gilkes and Gordon with Ringfeeder, the manufacturers of shaft fixing components.

2.6 Testing program

The test programme tested the integrity of the connection between the novel buckets and hub. It also indicated the extreme operating envelope of runaway speed and locked rotor torque. From these it was possible to estimate a safe maximum operating head.

The central disk, which represents the assembly of runner buckets, was clamped to a suitable rigid mounting. Drive was by hydraulic or pneumatic wrench bearing on the 110 mm A/F hexagon. A 60 mm A/F hexagon was machined on the central shaft - this was intended to augment the drive to the 110 mm hexagon if required.

Figure 2.5 - Assembly of test fixture, visualised by the solid modelling software

Figure 2.6 - Assembly prior to testing.

The following tests were performed at the testing facilities at Gilkes and Lancaster University.

Test Test Description Test results 1 Test hub to bucket security, fully assembled.

Secure inner ring to floor plate and rotate 110 mm hexagon with hydraulic wrench. Record the torque required to make the hub start to slip on the inner ring.

At maximum torque, 10,170 Nm (7,500 ft lbs) the hub root concept will transmit the required torque in a satisfactory manner, with a safety factor of at least 3.4 on the locked rotor torque.

2 Test hub to bucket security as in test 1, but with fully lubricated assembly. Secure inner ring to floor plate and rotate 110 mm hexagon with hydraulic wrench. Record the torque required to make the hub start to slip on the inner ring.

Slippage occurred at an applied torque of 5420 Nm. This would equate to an safety factor of 1.8 on locked rotor or 3.4 under normal operating service.

3 Test shaft to hub security by removing the screws securing the end plates to the hub halves and rotate the shaft by either hexagon. Record the torque to cause the shaft to just slip in the hub.

No movement was observed at the maximum applied torque of 10,170 Nm.

4 Repeat test 1 using two identical Ringfeeder 7013.1 fixings to examine the effect of any additional clamping effect due to the 7013.2 shaft fixing.

No movement was observed at the maximum applied torque of 10,170 Nm.

5 Repeat test 2 using two identical Ringfeeder 7013.1 fixings to examine the effect of any additional clamping effect due to the 7013.2 shaft fixing.

No movement was observed at the maximum applied torque of 10,170 Nm.

6 Repeat test 2 with a Ringfeeder 7013.2 removed and replaced with a Ringfeeder 7012 shaft fixing.

No movement was observed at the maximum applied torque of 10,170 Nm.

7 Repeat test 2 using two identical Ringfeeder 7012 fixings to examine the effect of using the different shaft fixings.

A barely perceptible movement was observed at the maximum applied torque of 10,170 Nm.

8 Test bucket bending deflection. Cut up the central disk to simulate the ring of buckets and fit a bucket. Apply a load to the bucket, measure the deflection to check the FEM calculations. A suitable load will not take the bucket out of the elastic region, but will give a deflection that is large enough to measure reliably.

The bucket remained wholly within the elastic region over the test load of 0 to 20kN. An acceptable non-linearity was observed which probably was caused by seating the bolted components by the testing machine.

9 Test bucket bending strength. Apply a load such that one of the ring segments eventually fails in bending or the clamp fails. Calculate the equivalent jet loading. This will indicate an ultimate head that can be safely applied to the runner.

The bucket failed at an applied force of 222 kN. The mode of failure was cracking in the bucket root near the neck radius. The test demonstrates that the bucket has sufficient static strength to withstand any conceivable working head.

10 Test bucket pull out force. Pull a ring segment in a radial direction until it or the clamp starts to plastically deform. Calculate the equivalent angular velocity. This will indicate the ultimate runaway speed and hence the maximum service speed.

The weld securing the attachment to the tensile tester failed at 238 kN. The bucket root was found to have just started to plastically deform at the bucket root.

11 Measure the force required to axially displace the hub. This will indicate the level of security of the runner upon the shaft for vertical shaft arrangements.

The runner slipped with an axial load of 297 kN, equivalent to a gravity load of 30.27 tonnes, and we see a safety factor of 50.

12 Assembly of full runner to assure production that

the runner can be built as designed and allow the craftsmen a final comment on design features and to modify the assembly and machining procedures if required.

It was found that there is a requirement to make the bucket roots rather simpler to manufacture. Gauging of the components has been raised as a problem by the industrial engineering department.

Determine the first resonant mode of buckets. It was found that resonance induced phenomena at bucket passing frequency is very unlikely. Resonance of the buckets at a harmonic of the bucket passing frequency will reduce the fatigue life considerably. The bucket passing frequency of this design is 60 Hz. for

13

four jets and the runner at its point of best efficiency. The resonant frequency of the bucket is 2,163 Hz. near the 36th harmonic of the bucket passing frequency it is very unlikely that resonance of the bucket will be a problem. If the runner speed is 1500 rpm, as designed, the bucket passing frequency is 50 Hz. for four-jet operation and the resonant frequency of the bucket is near the 43rd harmonic.

Figure 2.7 - Test 8, Bucket bending test to failure in the 100 tonne universal testing machine

2.7 Cost / Benefit & Conclusions The development of the BRB during the project has resulted in an innovation that has the potential to lead to substantial cost savings in runner manufacture. A UK patent has been applied for to protect the invention. This will be extended to cover each country where a competitor operates. A thorough testing program has shown the BRB will reduce manufacturing time and improve production control without compromising on operating performance.

Further Testing program Generally the test results were as expected. Some areas of further work are:-

• Test 10 reveals that the runaway speed may be the limiting factor in design, rather than the bending condition. A mild steel dummy bucket was pulled from the hub. Fortunately, the weld attaching the pulling mandrel to the major portion of the dummy runner failed, just as the bucket root entered the plastic region of the stress strain diagram. This allows us a chance to investigate the loosening that will occur prior to failure of the bucket root. The setscrews securing the hub halves were found to be slack when the rig was dismantled. The probable cause was the plastic yielding of the dummy bucket. The head corresponding to the observed force was calculated as 3149.7 metres, far beyond the operating window recommended for this design (normal maximum 250 metres). Therefore it can be concluded that the new runner will easily withstand the normal design runaway speed. This result indicates that higher operating heads may be possible though bucket erosion, cavitation and fatigue loading need to be taken into account.

• After Test 12, the Gilkes Industrial Engineering Department, who are responsible for the manufacture of the prototype turbine, expressed concern that it may be difficult to machine the radial faces on the bucket roots with adequate accuracy. An investigation is to be undertaken to assess the effects manufacturing tolerances on the bucket root, which is beyond the scope of this report.

Cost / Benefits assessment The new BRB design has the potential to lead to considerable savings in turbine manufacture while potentially leading to improvements in performance.

Manufacturing cost saving The new BRB design will dramatically reduce runner-manufacturing time. It was calculated that the first runner manufactured will save 55% of the cost of a conventional runner of the same capacity. This figure includes tooling that can be reused for subsequent machines of the same size. The estimated saving for subsequent units could be approximately 77% of the cost of conventional runner. Manufacturing time A significant part of finishing a conventional runner is the bucket grinding and polishing process. BRB technology will dramatically reduce this time by allowing easy access to all the surfaces to be machined and the new process eliminates all hand finishing operations. The lead-time to manufacture a runner is reduced considerably. If the runner is entirely of a new design and solid modelling of the buckets and tooling are required, a reduction in time of 68% can be expected. If new tooling only is required, a reduction in lead-time of 79% can be achieved. A replacement runner can be made to an existing design with an 89% saving in lead-time.

Hydraulic performance It is envisaged that hydraulic performance will improve due to greater control over the bucket surface geometry and finish. Fatigue considerations Historically, the problem with bolt-on buckets has been cyclic fatigue and corrosion exacerbating stress concentrations around the bolt leading to premature failure. The massive number of applied stress cycles, as high as high as 5x1010 over an operating lifetime, has been found to make fatigue failure the most likely failure mechanism. The FEM computer program has the capability to predict fatigue performance and reliable texts on the subject have been consulted, however the true test for the runner will be field trials.

Figure 2.8 - Assembled test runner

3 CONCENTRIC CASE PELTON

3.1 Project Concept The second part of the project concentrated on the concentric case manifold surrounding vertical shaft Pelton turbines. This was given the acronym CCVSP standing for Concentric Case Vertical Shaft Pelton. The aim was to engineer substantial cost savings into casing construction while producing viable designs for commercial manufacture. Computer aided design methods were used throughout the development process.

Model Parameters

A number of different turbine manifold configurations were conceptualised in the earlier stage of the project that may have been appropriate to the requirements of the project. Five final options were short listed to be modelled after analysis with the CFD (Computational Fluids Dynamics) program CFX 5.51.

The basic turbine parameters for the prototype turbine were selected after considering the casing dimensions from several existing vertical shaft pelton turbines in successful operation.

The selected basic parameters chosen for the ccvsp prototype can be summarised as:-

• All jets 82.6mm Diameter • Each jet 28.5 l/s (114 l/s total flow divided by 4 jets) • Internal wall roughness 0.05mm • Casing internal diameter 1323mm • Average Velocity in manifold 5m/s

3.2 Shortlisted Options The five options showing the most merit were modelled in depth using CFX 5.51 to establish their hydraulic performance. The preliminary results are shown below. (Fig 3.1)

Figure 3.1 Option Description CFX Contour Plot showing

Manifold Pressure (Pa)

1 • Rectangular manifold profile. Constant

depth, decreasing width. • Velocity designed to be 5 m/s adjacent

and through jets

2 • Rectangular manifold profile. Constant

width, decreasing depth. • Velocity designed to be 5 m/s adjacent

and through jets

No picture available – model did not proceed to 2nd assessment stage

3

• Round manifold profile. Reduction in

area immediately after jet branches. • Velocity designed to be 5 m/s adjacent

and through jets

3

• Toroid manifold. Constant manifold

diameter. • Velocity designed to be 5 m/s at inlet and

through jets

4

• Rectangular casing profile. Constant

depth, decreasing width. • Velocity designed to be 5 m/s adjacent to

jets on long leg

3.3 Shortlisted options Results / Evaluation

Headloss Comparison An indication of the hydraulic performance of each option is the hydraulic pressure loss from the model inlet to the end of the individual jets. The hydraulic headloss was calculated for each option by measuring the pressure at the inlet and jets using a tool within CFX. Figure 3.2 below show the pressure at each jet and hence the hydraulic headloss.

Figure 3.2 Headloss Comparison

3.4 Preferred Options for Modelling Options 1 & 5 were selected for further development based on their hydraulic performance following analysis with CFX 5.51 and their likely ease of manufacture. Construction of the test rig for physical modelling proceeded using a modelling scale of 0.5:1.0. The construction was undertaken by York Plastics of Yorkshire and the rig was delivered to Gilkes mid December 2002. The scale models operating parameters were :

• Maximum differential head = 5.0 metres • Maximum inlet flow = 20.2 l/sec. • Outlet flow = 5.05 l/sec. per nozzle

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

0 1 2 3 4 5

Jet Number

Hea

dlos

s (m

)

Option 1 Option 3 Option 4 Option 5

Testing Procedure

Figure 3.3 - The test rig set up at Gilbert Gilkes and Gordon’s test bay

The model was supplied with two 3-inch flexible pipe tails for connection to the Gilkes test system. The four outlet valves were threaded 2 inch BSP. The schematic below illustrates the test rig configuration.

Figure 3.4 - Schematic of test bay layout After connection to the external pipework, the following procedure to examine the flow distribution and headloss took place for each possible jet configuration:

1. Check the required jets are fully open

2. Check all the gauges are zeroed

3. Set pump speed to zero. Turn on pump and slowly increase the speed to give approximately the required flow rate

4. Wait for system flow to stabilise by observing pressure gauge fluctuation

5. Adjust backpressure by adjusting the jet valves until the pump flow rate is correct.

6. Once flow has stabilised, measure each individual jet’s flow and adjust the flow rate if required. Once each jet flow is correct, record the headloss from the manifold to the jet being considered in the run and note the flow conditions.

3.5 Hydraulic performance

Headloss results

The testing procedure for option 1 and option 5 was undertaken for every jet flow combination. Headloss from the inlet to the relative jet was recorded. Any notable flow occurrence during the test was noted. Table 3.5 below summarises the results.

Jet CFX ∆h

Miller ∆h

Average Measured

∆h

1 0.75m 0.78m 0.35m

2 0.71m 0.86m 0.34m

3 0.76m 0.87m 0.62m

4 0.72m

0.62m 0.46m

Table 3.5 - Comparison of the calculated and measured head loss through option 1 manifold, all four jets in operation, including CFX Pressure Contour plot.

In this case the values calculated by CFX and the Miller methodology are reasonably close. The values measured during testing however are consistently lower. This may have been caused by the proximity of the manifold pressure tapping to jet 1, which according to further CFD analysis, was subject to severe turbulence, which may have skewed these results. In fact this pressure tapping was recording a head gain rather than a head loss for some jet configurations.

Jet

CFX ∆h

Miller ∆h

Measured

∆h

1 0.47m 0.59m 0.80m

2 0.50m 0.76m 0.72m

3 0.70m 0.85m 0.29m

4 0.77m

1.13m 0.66m

Table 3.6 - Comparison of the calculated and measured head loss through option 5 manifold, all four jets in operation, including CFX Pressure Contour plot.

For option 5, the calculated head losses using CFX agree with the measured headlosses reasonably well. The exception is jet 3 where the calculated values appear rather high, though the geometry of the path leading to jet 3 is particularly irregular and difficult to calculate by theoretical means.

Test observations

Whilst testing the option 5 model, a vortex was observed to form occasionally in jet 2. This indicates that a low pressure region had developed though the effect on the hydraulic headloss was not noticeable.

Figure 3.7 - Vortex in jet 2 on the option 5 manifold

Figure 3.8 - showing mixing and consequent losses at jet 1 in the option 1 manifold - jets 1 and 4 open, 0.005 m3/sec.

In the region to the right of figure 3.9 there is no fluid flow; the green dye is in a stagnant region. It is thought that the region containing the dye is the cause of most of the losses observed. Figure 3.9 shows the stagnation and mixing from above.

Vortex

Figure 3.9 - View from above of mixing occurring at the manifold, jet 1 junction, jet 1 only open.

3.6 Prototype performance The purpose of the scale modelling was to help predict the hydraulic performance of the prototype manifold without the expense of full scale construction. The well understood laws of similarity mechanics were used to transpose the model performance to the prototype. The CFD software was used to estimate the prototype performance and the empirical hydraulic method was also included as a reference. Background to similarity mechanics Provided that a hydraulic machine is scaled uniformly – i.e. all geometric lengths are varied by a constant ratio, it is possible to calculate the performance of the prototype hydraulic machine using the principals of similarity mechanics. This method is based on well developed relationships between prototypes and their scaled equivalents. This allows the hydraulic performance of the prototype to be understood without the necessity of delving into complex fluid flow theory for its specific flow conditions.

The Euler scaling laws were able to be applied to the CCVSP test rig because of its turbulent fluid flows and corresponding high Reynolds numbers (>10^6). Equation 3.10 was derived for use in calculating the prototype casing results, which compares model and prototype unit flow, unit area and change in pressure.

Equation 3.10

Q 1

A 1 2P 1ρ

..

Q 2

A 2 2P 2ρ

..

Where: Q1 = Model unit flow Q2 = Prototype unit flow A1 = Model unit area A2 = Prototype unit area P1 = Model change in Pressure (Pa) P2 = Prototype change in Pressure

(Pa)

Empirical methods Texts are available (notably Miller) that offer a methodology to predict the pressure loss for a wide range of hydraulic components i.e. bends, tees, elbows. It was possible to separate the prototype manifold into its component parts and then add the individual losses together to form a picture of its overall hydraulic performance. An assessment of both model options was undertaken and are tabulated in tables 3.5 & 3.6 The limitations with empirical methods are inaccuracies due to Reynolds number variations, interpolation errors with specific flow cases and errors interpolating between non standard configurations – i.e. the model has components that fall in between two illustrated components and it was therefore necessary to interpolate between the two. Figure 3.11 below illustrates a typical table from Millers 1990.

Figure 3.11 Summary of Headlosses

To summarise, the headlosses in the prototype were calculated in two ways:-

1. Using the similarity mechanics equation 3.10, based on the model headlosses generated by CFX for scaling comparison.

2. A CFX model of the prototype was run using full-scale dimensions and a wall roughness of 0.05mm (equivalent to steel).

The results are presented in table 3.12 both option 1 and option 5 prototypes.

Based on Similarity Mechanics equation 3.10

CFX results

Jet Number Option 1 prototype

∆h (m)

Option 5 prototype

∆h (m)

Option 1 prototype

∆h (m)

Option 5 prototype

∆h (m)

1 1.56 1.60 1.49 1.51

2 1.48 0.97 1.34 0.89

3 1.59 1.00 1.42 0.90

4 1.50 1.46 1.42 1.35

Table 3.12 - Calculated head losses, all four jets in operation 3.7 Cost / Benefit & Conclusions

The new concentric case has the potential for significant project cost savings in three areas:- • Saving in powerhouse floor area required because of reduced manifold footprint. • Reduction in manifold construction costs because of greater simplicity.

An analysis was undertaken based on the layout of an existing 4 jet vertical shaft Pelton turbine installed by Gilkes in 1986. The existing installation was a very large capacity however, well beyond the expected operating range of the CCVSP, so the existing configuration was scaled down so a fair comparison could be made. The conventional manifolds construction cost, performance and footprint area were then compared to that of a option 1 type manifold of the same capacity. CFD was used to model the existing and theoretical cases. Cost comparison Prices were sought to manufacture the existing concentric case in the option 1 type design. The CCVSP was increased in capacity so that velocities in both manifolds were the same. Figure 3.13 below shows the existing Hatchet Creek manifold layout with the CCVSP (in blue) superimposed over it for comparison. The footprint area of the new manifold was 4.29m x 4.03m (17.3m²) compared to 5.1m x 5.1m (26.0m²) for the existing (excluding spear jet actuators in both cases). A typical powerhouse cost of £110k for a 100m² area was used in the analysis. Table 3.14 provides a cost comparison.

Figure 3.13 Powerhouse Construction Manifold Construction Existing Manifold £110,000 £35,000 New concentric case £107,000 £30,000 Saving (£k) £3,000 £5,000 % Saving 2.7% 14.3%

Table 3.14 The resulting overall saving in manufacturing and powerhouse cost is £8,000. Hydraulic performance

The following operating parameters were used in the hydraulic comparison between the existing and CCVSP manifolds: • All jet branch pipes 290mm diameter • Each jet 330l/s (1.32 m³/s total flow divided by 4 jets) • Internal wall roughness 0.05mm • Average velocity in manifold 5 m/s

Both manifolds were modelled using CFX to estimate headloss loss and hydraulic performance. The results are illustrated in table 3.15

Jet Existing manifold - ∆h (m)

New concentric case - ∆h (m)

1st 2.0m 0.94m

2nd 2.25m 0.99m

3rd 3.5m 0.96m

4th 3.5m 1.01m

Table 3.15

No onerous turbulence was observed in either model, probably due to the relatively low manifold velocities. The headloss experienced in the existing manifold is reasonably high in proportion to the schemes gross head – up to 2.5% of total energy, though this percentage loss would be less for higher operating heads.

The average reduction in jet headloss is therefore 1.84m (2.8m average headloss for existing less 0.975m average for CCVSP). To calculate the additional annual energy production gained from the saving in headloss achieved, an energy value of 5.5p/kW, and a plant load factor of 0.5 was used. This resulted in an additional £819 generation per annum for a typically sized scheme.

4 Conclusions Model testing It would have been more reassuring to have closer correlation between the theoretical and measured results for both options, however when taken in the context of the operating pressures of working manifolds (typically 150m+), the headlosses in table 5.3 above represent less than 1.0% of the total head within the system. Generally 1-3% is considered acceptable. The CFX results were verified by the Miller calculations and generally by the physical testing. CFX was established as an accurate tool that can determine manifold hydraulic performance. CCVSP design Both prototype manifolds had relatively low headlosses, though option 1 was considered to be better (lower cost, ease of construction and lower headloss.

Cost / Benefits The CCVSP design seems to have improved the turbine in three areas.

• A reduction in powerhouse construction costs of 2 to 3% may be possible due to the reduced footprint area required by the CCVSP.

• Manifold costs may be reduced by 14% • A modest reduction in headloss may be possible and may improve further if flow

baffles were incorporated into the branch pipe transitions where the major flow separation occurred. Baffles have successfully used where vortex formation has been a problem.