Investigation of flow through centrifugal pump impellers 4

CHAPTER EXPERIMENTAL WORK 4.1 General Introduction To verify numerical results, it is necessary to compare with experimental tests. The pump performance test was carried out using pump test facility with cold water. The speed, flow rate, suction pressure, discharge pressure, power and NPSHr (net positive suction head required) are measured during the test. Some of these data will be used to compare with the numerical results. In the pump performance test, the most commonly used driver is an electric motor. Due to fluctuations of power supply and load variation, the test speed usually fluctuates throughout the entire flow range. It is necessary to correct the pump performance of the test reading to a reference speed based on affinity law below. n1 Q = n2 Q2 ( n1 H ) = n2 H2 ( n1 P ) = n2 P2 (4.1) where n is the rotational speed in rpm, Q is the flow rate in m3/h, H is the total head in meter, P is the horsepower in kw, and subscripts and refer to speed and speed respectively. It is practically impossible for centrifugal pump to maintain a very steady flow rate and head without fluctuation in the reading during test. Therefore, it is imperative to take all the required readings simultaneously to ensure accuracy. A test method is thus developed to meet this requirement and to improve the efficiency so that the accurate test results can be obtained immediately at the end of the test. Figure 4.1 shows the schematic diagram of the test stand. 39 Chapter Experimental Work In order to choose a good pump test standard that is easily understood and applied by user, the most common pump test standard adopted by pump manufacturer today, that is, ISO 2548 (Centrifugal, mixed flow and axial pumps-Code for acceptance tests-Class C) was chosen as the standard of our pump test stand. This international standard is the first of a set dealing with acceptance tests of centrifugal, mixed flow and axial pumps; they correspond to three classes of tests A, B and C: class A is the most accurate and class C is the least accurate; the use of class A and B is restricted to special cases when there is a need to have the pump performance more precisely defined. Class C is commonly used for the mass produced pumps. The pump test stand discussed below is designed to meet ISO 2548. Figure 4.2 shows the test stand instrument arranged according to ISO 2548. For pump performance test, two assumptions are made: 1) Test liquid is the cold water, and the compressibility of test liquid is negligible. 2) Suction and delivery gauges are connected within times pipe diameter from the suction and delivery flanges of the pump. The frictional loss is negligible between pump flanges and gauge connection. 4.2 Theoretical Background 4.2.1 Total Head Calculation The most important parameter in the present study is the pump total head because it will be used for comparison with the computational value. Therefore, this section will introduce how to calculate the pump total head. Total head in meter at suction gauge position is given by 40 Chapter Experimental Work h1 + Ps V + s ρg 2g (4.2) and the total head in meter at discharge gauge position is given by h2 + Pd V + d ρg 2g (4.3) The pump total head in meter is obtained from the difference between Eq. (4.2) and Eq. (4.1) Pump total head H = h − h + − Vs Pd − P s V + d 2g ρg (4.4) where H is pump total head in meter, (h2-h1) is also refered to as difference in gauge height, P is pressure in Pa, V is velocity in m/s, h is the height in meter, and subscripts s and d refer to suction and discharge respectively. The velocity head is equal to the flow rate Q divided by the pipe cross-sectional area, hence: Vd = Q Ad Vs = Q As (4.5) where Vd and Vs are the velocity head of the delivery and suction point in m/s; Ad and As are the cross-sectional area of the delivery and suction pipe in m2. The suction and delivery pipe diameter are input into the computer during the test to calculate the pump total head based on the measured data. The suction and delivery pressure ( Ps and Pd ) are measured by means of digital pressure transmitter with output pressure indication in MPa and 4-20 mA output signal to central processor of the computer to compute the pump total head with the information of difference in gauge height. 41 Chapter Experimental Work 4.2.2 Calculation of Pump Power Input and Efficiency Assuming there is no loss in the power transmission with direct coupling drive between pump and motor, the pump power input can be written as Pp (pump power input) = Motor power output Motor power output = Pm (motor power input measured by the instruments) × η m Pp = Pm × η m Q×H 364 pu η = pp Pu = (4.6) where Pu is pump power output, η is pump efficiency, and η m is motor efficiency. 4.2.3 NPSH Test Net positive suction head (NPSH) refers to the energy of the test liquid at centerline of the pump above the vapour pressure of the test liquid. A valve throttling method is used to measure the NPSHr for the pump being tested. 3% head drop criterion, which is most commonly accepted internationally, is adopted. NPSH test is carried out by adjusting the pump capacity to the specified point. Suction valve is throttled to create vacuum at the pump suction, meanwhile the discharge valve is adjusted to maintain the capacity to be constant. The head will remain constant if the NPSHr is less than the NPSH available from the system. The test is repeated by further suction valve throttling until the head cannot be maintained at constant capacity. The NPSHr at 3% head drop is measured and calculated. The above test is repeated for different capacity so that a complete NPSHr curve for the pump being tested can be drawn. 42 Chapter Experimental Work 4.3 Experimental Facility 4.3.1 Description of Experimental Facility A fully computerised pump test facility is designed and built to obtain more accurate pump performance test data. This test facility is an open circuit and capable of testing pumps up to maximum flow rate of 2500 m3/h with maximum discharge pressure of 20 kg/cm2; the maximum power available is 500 kw. The flow is measured by a magnetic flow meter with accuracy within ± 0.25% of the readings. The output signal is 0~10,000 Hz. The suction and discharge pressure are measured by pressure transducers with output signal of 4~20 mA. The power measurement is done by Ampere meter, Volt meter and Power Factor meter. The test data are collected and processed by a personal computer to generate various test reports. Open and closed circuit test loops were considered during the test facility design. The advantage of closed loop is that the pump suction is always flooded and it eliminates the requirement for priming before starting the pump. It requires a large pressure vessel with proper cooling system to cool the water due to temperature rise during testing, a vacuum system is required to create a vacuum in the vessel in order to carry out the NPSHr test. The advantage of open test loop is simple construction and ease of operation. A simple priming system is required to prime the pump before the start up. To prevent turbulence that may affect the pump suction condition, two buffers were constructed in the underground water sump. The total volume of water in the sump should be large enough to maintain the temperature rise of the water below °C during the test. The velocity of the flow in the sump is maintained below 0.3 m/s to ensure that the suction condition will not have an adverse effect on the pump performance. 43 Chapter Experimental Work To run the test facility properly, special software has been developed to control its operation. The program size is about 10 GB, it is edited by Visual Basic, Borland C++ and Assembler language. The software can control the test facility in the following ways: (1) Data input including testing method selection (using electric motor performance curve or using dynamometer), pump type selection (centrifugal pump or submersible pump), units to be used in measurement selection (metric or U.S.A.) (2) Test run including taking readings for speed, power, suction and discharge pressure, voltage and current at each flow rate and converting these test data to the specified speed. (3) Generating all the test report including pump original report, pump performance report, pump performance curve, curve for the verification of testing standard, pump NPSHr data report and pump NPSHr curve. 4.3.2 Instrumentation The instruments for the pump test stand shown in Figure 4.2 are: • Danfoss MAG3100 digital output magnetic flow meter with flow calibrated in m3/h. A voltage is induced into a conductor which moves in a magnetic field. With the electromagnetic measuring principle, the flowing fluid is the moving conductor. The induced voltage is proportional to the flow velocity and is fed to the measuring amplifier by a pair of electrode. Using the pipe cross-sectional area, the flow is calculated. The transmitter converts the measured values coming from the sensors into standardised output signal. 44 Chapter Experimental Work These output signals are transferred to the computer for processing and generating the test reports. • Endress and Hauser pressure transmitter calibrated in MPa. The transmitter is based on the principle that the pressure acts directly on the separating diaphragm with a filling liquid transmitting the pressure to a resistance bridge. The bridge output voltage, which is proportional to the pressure, is measured, processed and converted to 4-20 mA output signal. The output signal is then fed into the computer to be processed. These instruments are selected based on market availability, and their accuracies are required to meet the International Standard ISO 2548 for centrifugal pump testing. The accuracies of various instruments are listed in Table 4.1 by referring to accuracy in percentage of reading. These instrument accuracies are higher than those used for general industrial applications and able to meet the permissible systematic errors of measuring instruments set out in ISO 2548. 4.4 Experimental Procedure The pump test was first started with the discharge valve closed to measure the shut off head and power. The test was continued by operating the discharge valve to vary the flow from shut- off to maximum flow. For each test, the rotational speed is set at 2900 rpm and 1450 rpm. About 14 groups of readings from shut down to 122%-158% design flow rate are taken for performance test, and groups of reading are taken for NPSHr test. The readings for pump capacity, rotation speed, power input, pump suction pressure and discharge 45 Chapter Experimental Work pressure are also taken at each flow rate. The performance test characteristics of design point are also recorded. For example, to test the fourth model pump M4 at rotational speeds of 2900 rpm and 1450 rpm, 14 groups of readings from shut down to 140% and 145% design flow rate were taken for performance test respectively, and groups of readings were taken for NPSHr test. The performance test characteristics of design point at rotational speed 2900 rpm are: Q = 270.84 m3/h, H = 154.21 m, BHP = 147.91 kW, η = 76.90%. The performance test characteristics of design point at rotational speed 1450 rpm are: Q = 181.20 m3/h, H = 36.56 m, BHP = 22.23 kW, η = 81.15%. All the above readings are recorded automatically in the computer. At each flow rate, readings of rotational speed, input power, suction pressure and discharge pressure were taken within 0.5 second, and the mean readings were obtained by averaging them. The pump performance curves are plotted based on the above test data. These curves are η -Q curve (efficiency verses capacity curve); H-Q curve (head verses capacity curve); P-Q curve (power verses capacity curve) and NPSHr-Q curve (NPSHr verses capacity curve). 4.5 Results and Discussion 4.5.1 Model Impellers Design Characteristics As discussed in Chapter 1, three of model impellers named M3, M4 and M5 will perform experimental test on the test stand described above and the experimental results will be used to compare with the numerical ones for validation. The design characteristics of these model impellers are summarised as follows: 46 Chapter Experimental Work For model impeller M3 at rotational speed of 2900 rpm, the design specifications are: Q = 27 m3/h, H = 96 m, and ns = 30. At 1450 rpm, the design specifications for M3 are: Q = 14 m3/h, H = 24 m, and ns = 30. The impeller inner diameter is 46 mm; outer diameter is 269 mm; and the outer width is mm with straight vanes. For model impeller M4 at rotational speed of 2900 rpm, the design specifications are: Q = 270 m3/h, H = 148 m, and ns = 68; at 1450 rpm, the design specifications for M4 are: Q = 180 m3/h, H = 36 m, and ns = 81. The impeller inner diameter is 106mm; outer diameter is 337mm; and the outer width is 21mm with twisted vanes. For model impeller M3 at 2900rpm, the design specifications are: Q=70m3/h, H=66m, and ns=64; at 1450rpm, the design specifications for M3 are: Q=45m3/h, H=14m, and ns=82. The impeller inner diameter is 70 mm; outer diameter is 221 mm; and the outer width is 14 mm with twisted vanes. 4.5.2 Experimental Results Discussions Figures 4.3-4.5 show the experimental H-Q curves for model impellers M3, M4 and M5 at two rotational speeds of 2900 rpm and 1450 rpm. The converted curves using similarity law are also drawn on each graph to make comparison with the experimental data. It is found that the pump head H in column of water decreases as the flow rate Q through the pump impeller increases and there is no linear relation between H and Q. For the actual pump head H, we have the following relation: H = H T − hloss H T = a − bQ hloss = cQ (4.7) where H T is theoretical pump head; hloss is head losses; a, b and c are constant. 47 Chapter Experimental Work It can be seen that although theoretical pump head H T decreases linearly as flow rate Q increases, head losses hloss doesn’t follow linear relation with Q. If only two main types of energy losses that occur in the pump are considered, friction loss at the blades and passage surfaces and shock loss due to the mismatch between the blade angle and the inlet flow direction, and neglect other types of losses, head losses hloss will be propotional to Q . From the above figures, a close relation between experimental curve and predicted curve using similarity law (4.1) is also shown on each graph; this shows the similarity law is valid for all the test curve, with this, all the test data given can be formulated using similarity law given in Eq. (4.1). This also proves the pump test facility is capable of performing to the required standard. Figures 4.6-4.8 show the experimental η -Q curves for model impellers M3, M4 and M5 at two rotational speeds. Several observations can be made from these figures. Firstly, it is found that the efficiency of the pump η increases as the flow rate Q increases. However, the efficiency starts to decrease as the flow rate increases further. This can be verified from relationship (4.6). The water power of the pump (Pu) will initially increase more than the motor power output (Pp) when the flow rate (Q) is increased. Thus, the pump efficiency ( η ) will also be increased, however, after it reaches its maximum efficiency, the shock loss, frictional losses, circulation loss and other losses that occur in the pump will become more prominent and increase more rapidly than the head supplied. Hence, the power output of the pump (Pu) will decrease gradually and the motor power of the pump (Pp) will increase continuously. As a result, the pump efficiency begins to drop with increasing flow rate. The experimental results verify this 48 Chapter Experimental Work trend. Secondly, from the figures, it can be seen that the efficiencies of model impellers M4 and M5 are much higher than that of model impeller M3. For example, the experimental results show that the maximum pump efficiency for M4 and M5 are 77% and 74% respectively at n = 2900 rpm, whereas the maximum pump efficiency for M3 is only about 55%. Therefore, it can be concluded that the efficiency of twisted vane pump will be higher than that of straight vane pump. Thirdly, the experimental results also show that the maximum efficiency occurs near pump design point, hence, impeller design is acceptable. Figures 4.9-4.11 show the experimental P-Q curves for model impellers M3, M4 and M5 at two rotational speeds of 2900 rpm and 1450 rpm. The converted curves using similarity law are also drawn on each graph to make comparison with the experimental data. It is seen that the shaft power of the pump increases linearly as the flow rate increases. From the similarity laws, the shaft power of pump is directly proportional to the third order of the rotational pump speed (n). Hence, for pump running at higher rotational speed, more shaft power will be required to maintain constant rotational speed. After comparing the experimental curves and the predicted curves at n = 2900 rpm and 1450 rpm, it is found that these curves are quite close and identical. This shows that when shaft power of a pump operating at a particular speed is available, the shaft power of the pump operating at another speed can be accurately predicted without really doing the whole experiment again, by using the similarity law method. Hence, the similarity law is an efficient and simple method for calculation of the pump characteristics at different speeds. Thus, the similarity law is also verified. Figures 4.12-4.14 show NPSHr test data for model impellers M3, M4 and M5 at two rotational speeds. C Q is non-dimensional flow coefficient, its definition is: 49 Chapter CQ = Experimental Work Q nD where Q is flow rate in m/s; n is rotational speed in rps; D is diameter of pump in m. NPSH value, which is demonstrable by an industrial grade test, provides a commercially acceptable basis for the quantitative assessment of the total NPSH necessary to suppress unacceptable levels of cavitation. And NPSHr is usually defined as the NPSH required at the pump inlet for operation free from unacceptable deficiencies caused by cavitation. It is found that NPSHr value increases as the flow rate increases. This can be proved by the definition of NPSH which is: NPSH= { total inlet head } – { head equivalent of vapour pressure} This gives: NPSH = [hs1 − hF + Pvap Ps V12 + ]− ρg g ρg (4.8) where hs1 is the static head at the pump inlet; hF is the head loss at the pump inlet; Ps is pressure at the pump inlet; V1 is the mean velocity of liquid at the pump inlet; Pvap is vapour pressure; ρ is liquid density. From Eq. (4.8), it is clear that NPSH value will be increased as the mean velocity at the pump inlet ( V1 ) increases. Since V1 is proportional to the flow rate (Q), that means NPSH value will depend on the inlet flow rate. It is also seen from Eq. (4.8) that NPSH value can be increased by increasing hs1 and Ps values. This can be achieved by adjusting the pump installation height. Improving pump suction condition can reduce the head loss at the pump inlet and thus suppress the cavitation. On the other hand, although it is reasonable to guess that NPSH value may also depend on rotational speed, we cannot get any conclusion from our test data. For model impeller M4 and M5, NPSH value at 2900 rpm is higher than NPSH value at 1450 rpm. But for model impeller M3, the result 50 Chapter Experimental Work is on the contrary. This suggests that accurate NPSHr test data can be obtained by experiment only, similarity law can not be used to predict NPSH value as for pump head and shaft power of the pump. 51 Chapter Experimental Work Power meter Pressure transducer Flow meter Figure 4.1 The schematic diagram of the test stand Flow meter Pd Ps valve pump Figure 4.2 The instruments arrangement of test stand 52 Chapter Experimental Work 100 25 80 20 60 H (m) 30 H (m) 120 Test data n=2900rpm 40 20 40 Converted data from n=2900rpm 0 Test data n=1450rpm 10 Converted data from n=1450rpm 20 15 60 10 20 30 Q (m^3/hr) Q (m^3/hr) (a) n = 2900 rpm (b) n = 1450 rpm 50 180 160 140 120 100 80 60 40 20 40 H (m) H (m) Figure 4.3 Comparison of test data and converted data for model impeller M3 Test data n=2900rpm Converted data from n=1450rpm 200 400 30 20 Test data n=1450rpm 10 Converted data from n=2900rpm 0 600 100 200 Q (m^3/hr) Q (m^3/hr) (a) n = 2900 rpm (b) n = 1450 rpm 300 80 20 70 60 15 50 40 H (m) H (m) Figure 4.4 Comparison of test data and converted data for model impeller M4 Test data n=2900rpm 30 20 50 100 Q (m^3/hr) (a) n = 2900 rpm Test data n=1450rpm Converted data from n=1450rpm 10 10 150 Converted data from n=2900rpm 0 20 40 60 80 Q (m^3/hr) (b) n = 1450 rpm Figure 4.5 Comparison of test data and converted data for model impeller M5 53 Chapter η (%) Experimental Work 60 60 50 50 40 η (%) 30 20 Test Data 10 Trendline (Test Data) 0 10 20 30 40 30 20 Test Data 10 Trendline (Test Data) 40 Q (m^3/hr) 10 20 30 Q (m^3/hr) (a) n = 2900 rpm (b) n = 1450 rpm Figure 4.6 Experimental η -Q curves for model impeller M3 100 90 80 70 60 η (%) 50 40 30 20 10 η (%) Test Data Trendline (Test Data) 100 200 300 400 80 60 40 Test Data 20 Trendline (Test Data) 0 500 100 200 300 Q (m^3/hr) Q (m^3/hr) (a) n = 2900 rpm (b) n = 1450 rpm Figure 4.7 Experimental η -Q curves for model impeller M4 90 80 70 60 η (%) 50 40 30 20 10 70 60 50 η (%) Test Data Trendline (Test Data) 40 30 Test Data 20 Trendline (Test Data) 10 50 100 Q (m^3/hr) (a) n = 2900 rpm 150 20 40 60 80 Q (m^3/hr) (b) n = 1450 rpm Figure 4.8 Experimental η -Q curves for model impeller M5 54 Chapter Experimental Work 20 2.5 2.0 P (Kw) P (Kw) 15 10 Test data n=2900rpm Converted data from n=1450rpm 1.5 1.0 Test data n=1450rpm 0.5 Converted data from n=2900rpm 0.0 20 40 60 10 20 30 Q (m^3/hr) Q (m^3/hr) (a) n = 2900 rpm (b) n = 1450 rpm 250 30 200 25 150 100 P (Kw) P (Kw) Figure 4.9 Comparison of test data and converted data for model impeller M3 Test data n=2900rpm 50 20 15 Test data n=1450rpm 10 Converted data from n=1450rpm Converted data from n=2900rpm 0 200 400 600 100 Q (m^3/hr) 200 300 Q (m^3/hr) (a) n = 2900 rpm (b) n = 1450 rpm Figure 4.10 Comparison of test data and converted data for model impeller M4 30 20 P (Kw) P (Kw) 25 15 Test data n=2900rpm 10 Test data n=1450rpm Converted data from n=1450rpm Converted data from n=2900rpm 0 50 100 Q (m^3/hr) (a) n = 2900 rpm 150 20 40 Q (m^3/hr) 60 80 (b) n = 1450 rpm Figure 4.11 Comparison of test data and converted data for model impeller M5 55 Chapter Experimental Work NPSHr (m) n=2900 rpm Trendline (n=2900 rpm) n=1450 rpm Trendline (n=1450 rpm) 0 0.5 C Q (× 10 1.5 −3 ) Figure 4.12 NPSHr test data for model impeller M3 at two rotational speeds 10 n=2900 rpm NPSHr (m) n=1450 rpm Trendline (n=2900 rpm) Trendline (n=1450 rpm) 0 10 15 C Q (× 10 20 −3 25 30 ) Figure 4.13 NPSHr test data for model impeller M4 at two rotational speeds n=2900 rpm NPSHr (m) n=1450 rpm Trendline (n=2900 rpm) Trendline (n=1450 rpm) 0 C Q (× 10 − ) Figure 4.14 NPSHr test data for model impeller M5 at two rotational speeds 56 Chapter Experimental Work Table 4.1 Accuracies of test stand instruments Instrument Accuracy Magnetic flow meter ± 0.25% Pressure transmitter ± 0.25% Power meter ± 0.5% Amp meter ± 0.5% Volt meter ± 0.5% Speed sensor ± 0.5% 57 [...]... Figure 4. 1 The schematic diagram of the test stand Flow meter Pd Ps valve pump Figure 4. 2 The instruments arrangement of test stand 52 Chapter 4 Experimental Work 100 25 80 20 60 H (m) 30 H (m) 120 Test data n=2900rpm 40 20 40 Converted data from n=2900rpm 5 0 0 0 Test data n= 145 0rpm 10 Converted data from n= 145 0rpm 20 15 0 60 10 20 30 Q (m^3/hr) Q (m^3/hr) (a) n = 2900 rpm (b) n = 145 0 rpm 50 180 160 140 ... 100 80 60 40 20 0 40 H (m) H (m) Figure 4. 3 Comparison of test data and converted data for model impeller M3 Converted data from n= 145 0rpm 200 40 0 20 Test data n= 145 0rpm 10 Test data n=2900rpm 0 30 Converted data from n=2900rpm 0 0 600 100 200 Q (m^3/hr) (a) n = 2900 rpm 300 Q (m^3/hr) (b) n = 145 0 rpm Figure 4. 4 Comparison of test data and converted data for model impeller M4 20 70 60 15 50 40 H (m)... Trendline (n= 145 0 rpm) 5 4 3 2 1 0 0 5 10 15 C Q (× 10 20 −3 25 30 ) Figure 4. 13 NPSHr test data for model impeller M4 at two rotational speeds 6 n=2900 rpm NPSHr (m) 5 n= 145 0 rpm Trendline (n=2900 rpm) 4 Trendline (n= 145 0 rpm) 3 2 1 0 0 1 2 3 4 5 6 C Q (× 10 − 3 ) Figure 4. 14 NPSHr test data for model impeller M5 at two rotational speeds 56 Chapter 4 Experimental Work Table 4. 1 Accuracies of test stand... rpm 150 0 20 40 60 80 Q (m^3/hr) (b) n = 145 0 rpm Figure 4. 8 Experimental η -Q curves for model impeller M5 54 Chapter 4 Experimental Work 20 2.5 2.0 P (Kw) P (Kw) 15 10 0 1.0 Test data n= 145 0rpm 0.5 Test data n=2900rpm Converted data from n= 145 0rpm 5 1.5 Converted data from n=2900rpm 0.0 0 20 40 60 0 10 20 30 Q (m^3/hr) Q (m^3/hr) (a) n = 2900 rpm (b) n = 145 0 rpm Figure 4. 9 Comparison of test data... be concluded that the efficiency of twisted vane pump will be higher than that of straight vane pump Thirdly, the experimental results also show that the maximum efficiency occurs near pump design point, hence, impeller design is acceptable Figures 4. 9 -4. 11 show the experimental P-Q curves for model impellers M3, M4 and M5 at two rotational speeds of 2900 rpm and 145 0 rpm The converted curves using... (b) n = 145 0 rpm Figure 4. 6 Experimental η -Q curves for model impeller M3 100 90 80 70 60 η (%) 50 40 30 20 10 0 η (%) 80 60 40 Trendline (Test Data) 0 100 200 300 40 0 Test Data 20 Test Data Trendline (Test Data) 0 0 500 100 200 300 Q (m^3/hr) Q (m^3/hr) (a) n = 2900 rpm (b) n = 145 0 rpm Figure 4. 7 Experimental η -Q curves for model impeller M4 90 80 70 60 η (%) 50 40 30 20 10 0 70 60 50 η (%) 40 Trendline... different speeds Thus, the similarity law is also verified Figures 4. 12 -4. 14 show NPSHr test data for model impellers M3, M4 and M5 at two rotational speeds C Q is non-dimensional flow coefficient, its definition is: 49 Chapter 4 CQ = Experimental Work Q nD 3 where Q is flow rate in m/s; n is rotational speed in rps; D is diameter of pump in m NPSH value, which is demonstrable by an industrial grade test,... n=2900rpm 50 20 15 Test data n= 145 0rpm 10 Converted data from n= 145 0rpm 5 0 Converted data from n=2900rpm 0 0 200 40 0 600 0 100 Q (m^3/hr) 200 300 Q (m^3/hr) (a) n = 2900 rpm (b) n = 145 0 rpm Figure 4. 10 Comparison of test data and converted data for model impeller M4 30 4 3 20 P (Kw) P (Kw) 25 15 Test data n=2900rpm 10 5 Test data n= 145 0rpm 1 Converted data from n= 145 0rpm 0 2 Converted data from n=2900rpm... rpm 150 0 20 40 Q (m^3/hr) 60 80 (b) n = 145 0 rpm Figure 4. 11 Comparison of test data and converted data for model impeller M5 55 Chapter 4 Experimental Work 7 6 NPSHr (m) 5 4 3 2 n=2900 rpm 1 Trendline (n=2900 rpm) n= 145 0 rpm Trendline (n= 145 0 rpm) 0 0 0.5 1 C Q (× 10 1.5 −3 2 ) Figure 4. 12 NPSHr test data for model impeller M3 at two rotational speeds 10 n=2900 rpm NPSHr (m) 9 8 n= 145 0 rpm Trendline... 2900 rpm Test data n= 145 0rpm 5 Converted data from n= 145 0rpm 10 0 10 150 Converted data from n=2900rpm 0 0 20 40 60 80 Q (m^3/hr) (b) n = 145 0 rpm Figure 4. 5 Comparison of test data and converted data for model impeller M5 53 Chapter 4 Experimental Work 60 η (%) 60 50 50 40 η (%) 30 20 Test Data 10 Trendline (Test Data) 10 20 30 30 20 Test Data 10 0 0 40 Trendline (Test Data) 0 40 0 Q (m^3/hr) 10 20 . common pump test standard adopted by pump manufacturer today, that is, ISO 2 548 (Centrifugal, mixed flow and axial pumps-Code for acceptance tests-Class C) was chosen as the standard of our pump. g V g P h d d 2 2 2 + + ρ (4. 3) The pump total head in meter is obtained from the difference between Eq. (4. 2) and Eq. (4. 1) Pump total head H g V V g P P h h s d s d 2 2 2 1 2 − + − + − = ρ (4. 4) where H is pump. each flow rate. The performance test characteristics of design point are also recorded. For example, to test the fourth model pump M4 at rotational speeds of 2900 rpm and 145 0 rpm, 14 groups of