Experimental versus EPANET Simulation of Variable Speed Driven

Available online at www.sciencedirect.com

ScienceDirect Energy Procedia 112 (2017) 100 – 107

Sustainable Solutions for Energy and Environment, EENVIRO 2016, 26-28 October 2016, Bucharest, Romania

Experimental versus EPANET Simulation of Variable Speed Driven Pumps Operation Georgiana Duncaa, Vlad-Florin Piraianua, Razvan Romana, Petre-Ovidiu Ciuca,b, SandaCarmen Georgescua,* a

Power Engineering Faculty, University “Politehnica” of Bucharest, 313 Spl. Independentei, Bucharest 060042, Romania b Multigama Service S.R.L., 9 Sapte Drumuri, Bucharest 031646, Romania

Abstract This study points on the operation of a pumping station test rig equipped with 3 vertical centrifugal pumps, with variable speed, coupled in parallel. We investigated the pumps operation by imposing reference levels for the pumping head, then attaining the requested duty points through rotational speed control. On this test rig, the pumped flow rate is controlled by a built-in software, where the pumping scheduling algorithm relies on a sensor-less control method, based on the pump performance curve and affinity laws. The setup allows the measurement of the pumps head (displayed on rig's control panel) and rotational speed of each motor. According to the pumping schedule, when a number of n pumps are operating (where n 2 or n 3 ), then (n  1) pumps run at nominal speed, while the remaining pump runs at a speed lower than, or equal to the nominal speed. We built in EPANET the numerical model of this test rig and calibrate the model upon experimental data, using simple control statements for the discharge valve and speed patterns for pumps. Due to the calibration, a perfect match resulted between experimental readings and numerical duty points. Further, we built an alternative numerical model, which allowed analysing additional duty points, then we plot them on a chart, together with the experimental duty points: all points fit into the operation areas attached to the pumping schedule. The proposed approach proved to be useful to study further the pumps operation, with a minimal experimental effort. © 2017 2017Published The Authors. Published by Elsevier Ltd. © by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license Peer-review under responsibility of the organizing committee of the international conference on Sustainable Solutions for Energy (http://creativecommons.org/licenses/by-nc-nd/4.0/). Peer-review under responsibility of the organizing committee of the international conference on Sustainable Solutions for Energy 2016. and Environment and Environment 2016 Keywords: variable speed driven pump; pumping scheduling; duty point; affinity laws; EPANET

* Corresponding author. Tel.: +4-072-362-4418; fax: +4-021-318-1015. E-mail address: [email protected]

1876-6102 © 2017 Published by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license (http://creativecommons.org/licenses/by-nc-nd/4.0/). Peer-review under responsibility of the organizing committee of the international conference on Sustainable Solutions for Energy and Environment 2016 doi:10.1016/j.egypro.2017.03.1070

Georgiana Dunca et al. / Energy Procedia 112 (2017) 100 – 107

1. Introduction In a hydraulic system, the operation of centrifugal pumps under variable water demand can be conducted by 4 methods [1]÷[4]: throttling control (using the discharge valve), bypassing control (reducing the flow output to consumers using a by-pass pipe, to redirect part of the pumped flow back to the suction side), rotational speed control (using a frequency converter), and on-off control (where the pump is either running or stopped, to keep the pressure in a tank between imposed limits). Regarding the wire to water efficiency of the process, the pump operation control was assessed by Georgescu et al [1] through a numerical model built in EPANET, able to simulate the hydraulic system operation under all 4 control methods. In this paper, we will focus on the rotational speed control method, which proved to be an energy efficient control method [5]÷[11]. Our study points on the operation of a pumping station test rig found in a Hydraulic Machinery Laboratory at the University "Politehnica" of Bucharest. The studied test rig [12] (designed by DP-Pumps as booster system for the domestic market) is a closed loop hydraulic system fed from a water tank; the rig is equipped with 3 vertical centrifugal pumps coupled in parallel, with suction and discharge connections in-line (figures 1 and 2). The pumps, labelled as P1, P2 and P3 in figure 2, are identical, their rated duty point being defined by 1.78 m 3/h ( 0.494 ˜ 103 m3/s) flow rate, 18 m head, 53.4% efficiency and 0.16 kW power; the nominal speed is n0 3430 rpm, at the frequency f 0 60 Hz. The pumps are driven with variable rotational speed, through frequency converters. The rotational speed nk d n0 of each motor ( k 1y 3 , for pumps P1÷P3) can be derived from the frequency value f k d 60 Hz, which is readable on the LCD display module attached to each frequency converter (figure 1). The frequency converters have an analogue unified output of 420 mA and an accuracy of r0.8% of full scale, so the readings (frequency values) can be affected by r0.4 Hz.

Fig. 1. Experimental rig: overview and detail on the frequency converters mounted inside the control panel.

Fig. 2. Experimental rig: pumping units (labelled from P1 to P3), discharge control valve (TCV) and other components

On the above experimental rig, we investigated the pumps operation by setting a reference level for the pumping head H attached to the parallel pumps operation point, then gradually open/close the Throttle Control Valve  TCV (figure 2), to attain different duty points through rotational speed control. On this rig, the pumped flow rate is controlled by a built-in software, where the pumping scheduling algorithm relies on a sensor-less control method, based on the pump performance curve (head versus flow rate curve) H k H k Qk , and affinity laws written as:

101

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Georgiana Dunca et al. / Energy Procedia 112 (2017) 100 – 107

Hk H0

nk

n0 2

rk2 and Qk Q0

nk n0

rk ,

(1)

where rk nk n0 f k f 0 is the speed ratio; the subscript 0 refers to hydraulic parameters at nominal speed n0 (head H 0 and pumped flow rate Q0 ), while k 1y 3 , for pumps P1÷P3, refers to the head H k and flow rate Qk at the rotational speed nk d n0 . For the centrifugal pumps mounted on the studied rig, the performance curves and efficiency curves governing the pumps operation at different rotational speeds are written based on affinity laws (1) [1]÷[3], [13] as following: 2 2 ­ ° H k H k Qk 31.62 rk  1.36 Qk , ® 2 ° ¯K k K k Qk 45.75 Qk rk  9.8 Qk rk .

(2)

where k 1y 3 , for pumps P1÷P3. For a duty point defined by the head H k , pumped flow rate Qk and pump efficiency K k , the pump's power can be computed as: Pk UgQk H k Kk , where U is the water density and g is the gravity. The total pumped flow rate Q summarizes all Qk values at a given time moment. Within the present study, we assume that the pump head H k , attached to pump's duty point, equals the parallel coupling pumping head H . The H values, in meters, are obtained based on the differential pressure 'p recordings yielded by 2 pressure transducers, mounted on the discharge and suction pipes (figure 2); the absolute pressure transducers (in the range 06 bar) have an accuracy of r0.5% of full scale. The values of 'p are displayed in bar on the rig's control panel, and are related to the pumping head as: H k # H 'p ( Ug ) . Based on the gathered experimental data, we built and calibrated in EPANET the numerical model of this test rig. Then, we built and validated in EPANET an alternative numerical model, by replacing the TCV type valve with a General Purpose Valve  GPV [3], [14]. Thus, we continued to study the pumps operation within the above hydraulic system, without additional experimental effort. 2. Experimental results The studied pumping station test rig operates according to a pumping schedule built on the following rules: x when a single pump is open, it runs at a speed lower than or equal to the nominal speed; due to reliability reasons, the order in which a pump is put in isolated operation mode varies (the built-in control software selects the working pump upon the total duration of the operation of those pumps, as well as upon the number and frequency of their starts and stops); x when a number of n pumps are operating in parallel, where n 2 or n 3 , then (n  1) pumps run at nominal speed, while the remaining pump runs at a speed lower than or equal to the nominal speed; here also, due to reliability reasons, the order in which a pump is put in parallel-coupled operation mode varies. Several experimental campaigns were conducted on the above test rig, by setting different reference levels for the pumping head H . During the experiments, the above reference level had a qualitative role [12], the measured pumping head H varying within a narrow range around the reference head value. For each reference head value, the flow rate was increased gradually (using the TCV), from zero up to a maximum value (when increasing the pumped flow rate, the pumps are opened one by one, through rotational speed control). After reaching the parallel pumps operation point where all pumps run at nominal speed, the flow rate was gradually decreased down to zero (the pumps being closed one by one, according to the speed control algorithm). The experimental investigations were concluded by 63 parallel pumps operation points, for which some measured and computed parameters are summarized in table 1, namely: the values of the pumping head H (in meters) and frequencies f k d 60 Hz, the speed ratios rk f k f 0 f k 60 , where k 1y 3 , and the total pumped flow rate Q (in m3/h). For the conducted experiments, the pumping head H varied in a range bordered by the upper limit H max 24.7 m and the lower limit H min 18.1 m. According to the rotational speed control method, the pumps' speed varied between the minimum value nmin 0.792n0 ( rmin 0.792 in table 1) and the nominal speed n0 .

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Georgiana Dunca et al. / Energy Procedia 112 (2017) 100 – 107 Table 1. Measured experimental data: pumping head H in meters, frequency values f k d 60 Hz; and computed data: speed ratios rk , where k 1y 3 for pumps P1÷P3, total pumped flow rate Q in m3/h and loss coefficient values ] at the TCV. No.

H [m]

f1 [Hz]

f2 [Hz]

f3 [Hz]

r1 = f1/f0

r2 = f2/f0

r3 = f3/f0

Q [m3/h]

] []

1

23.3

0

52

0

0

0.867

0

0.575

28932

2

23.5

0

52.5

0

0

0.875

0

0.722

24200

3

23.9

0

60

52.4

0

1

0.873

2.782

1325

4

24.0

0

60

54.6

0

1

0.910

3.634

825

5

22.4

0

60

60

0

1

1

5.207

355

6

24.0

52.5

60

60

0.875

1

1

5.126

364

7

23.5

0

0

52.2

0

0

0.870

0.564

25000

8

19.5

47.5

0

0

0.792

0

0

0.483

41032

9

19.4

52.4

0

0

0.873

0

0

1.862

3140

10

18.6

60

0

0

1

0

0

3.094

865

11

19.8

60

48.6

0

1

0.810

0

3.782

609

12

19.7

60

51.5

0

1

0.858

0

4.587

451

13

20.0

60

58.4

0

1

0.973

0

5.629

264

14

19.7

60

60

0

1

1

0

5.921

231

15

20.1

60

60

48

1

1

0.800

6.138

276

16

18.1

60

60

60

1

1

1

9.459

73

17

24.3

60

55.6

60

1

0.927

1

3.768

835

18

21.7

60

60

60

1

1

1

5.402

330

19

24.3

60

53

60

1

0.883

1

5.163

311

20

24.2

60

0

56.5

1

0

0.942

4.016

550

21

24.4

60

0

53

1

0

0.883

2.752

880

22

23.6

55.4

0

0

0.923

0

0

1.571

3450

23

23.9

53

0

0

0.883

0

0

0.754

18800

24

23.8

55.4

0

0

0.923

0

0

1.524

4970

25

24.1

60

0

52.5

1

0

0.875

2.635

1399

26

24.2

60

0

52.5

1

0

0.875

2.417

1800

27

23.8

60

0

54.7

1

0

0.912

3.748

758

28

24.0

60

0

59.2

1

0

0.987

4.600

485

29

24.5

60

54.2

60

1

0.903

1

5.555

350

30

22.4

60

60

60

1

1

1

7.811

157

31

23.8

0

52.5

0

0

0.875

0

0.548

31850

32

23.9

0

53

0

0

0.883

0

0.754

16600

33

23.6

0

52

0

0

0.867

0

0.332

99999

34

23.4

0

55.3

0

0

0.922

0

1.595

4140

35

24.3

53

60

0

0.883

1

0

2.843

1230

36

24.1

53.6

60

0

0.893

1

0

3.265

981

37

23.9

56.4

60

0

0.940

1

0

4.106

580

38

24.4

57.2

60

0

0.953

1

0

4.090

638

39

24.5

60

60

54.9

1

1

0.915

5.781

315

40

23.4

60

60

59.3

1

1

0.988

7.263

179

f k f0

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Georgiana Dunca et al. / Energy Procedia 112 (2017) 100 – 107 41

22.6

60

60

60

1

1

1

7.726

152

42

22.5

60

60

60

1

1

1

7.769

148

43

24.5

60

60

59.4

1

1

0.990

6.761

194

44

24.5

60

60

53.6

1

1

0.893

5.311

283

45

24.7

60

56.2

0

1

0.937

0

3.751

248

46

23.5

56.4

0

0

0.940

0

0

1.807

2600

47

23.9

53.8

0

0

0.897

0

0

1.058

8552

48

24.0

53.5

0

0

0.892

0

0

0.916

14600

49

23.5

59.4

0

0

0.990

0

0

2.347

1815

50

24.4

60

0

53

1

0

0.883

2.752

1520

51

23.9

60

0

59.4

1

0

0.990

4.666

460

52

24.5

60

53.3

60

1

0.888

1

5.153

382

53

24.2

60

54.7

60

1

0.912

1

5.908

302

54

23.5

60

59.9

60

1

0.998

1

7.315

176

55

22.1

60

60

60

1

1

1

7.937

152

56

23.6

0

0

52

0

0

0.867

0.332

86745

57

23.7

0

0

52.5

0

0

0.875

0.612

28980

58

23.2

0

0

56

0

0

0.933

1.787

3480

59

24.4

52.2

0

60

0.870

0

1

2.304

1960

60

24.2

54.4

0

60

0.907

0

1

3.484

910

61

23.9

60

0

60

1

0

1

4.765

498

62

23.1

60

60

60

1

1

1

7.509

153

63

24.5

60

55.6

60

1

0.927

1

5.973

268

3. Numerical models built in EPANET and numerical results We built in EPANET the numerical model of the pumping station test rig described in Section 1. The hydraulic system map is presented in figure 3, where all main components are labelled: water tank (total head of 0.6 m, kept constant), pumps P1÷P3, Throttle Control Valve  TCV on the main discharge pipe (where different values of the minor loss coefficient ] can be provided, upon the flow rate [3], [14]). The diameter of the main discharge and suction pipes is of 40 mm; pipes' length is displayed in meters on the map; the roughness of pipes wall was set equal to 0.1 mm. For each pump, we set in EPANET the performance curve and the efficiency curve described by the second order polynomials (2) at the nominal speed, namely for rk 1 , where k 1y 3 . Our first goal was to use that TCV-numerical model as a tool, tuned upon experimental readings, to obtain a proper curve ] ] Q . So, we calibrated the TCV-numerical model upon experimental data, using speed patterns for pumps and simple control statements for the discharge valve (TCV in figure 2). Within the Hydraulics Options of EPANET [3], [14], the flow units were set as CMH (m3/h) and the Darcy-Weisbach formula was selected for the headloss. According to the speed ratio values depicted from table 1, and considering that a full simulation develops over 63 time moments t (for the 63 experimental tests), a time dependent speed pattern was implemented in EPANET for each pump, namely rk rk t , where k 1y 3 for pumps P1÷P3. Knowing that at each time moment t , the total pumped flow rate Qt must fit the values from table 1, we determined by trial and error the appropriate values of the minor loss coefficient ] ] t to be set at the valve labelled as TCV in figure 3; the resulting ] values are inserted on the last column of table 1. Within the Simple Controls Editor of EPANET [3], [14], 63 simple if-statement rules were inserted, e.g. valve TCV 28932 at time 1, ..., valve TCV 268 at time 63, where ] 1 28932 ; ] 63 268 (table 1). As expected, due to the calibration, a perfect match resulted between experimental readings and numerical duty points. To exemplify the numerical results, we present in figure 4 the flow rate distribution through the hydraulic

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Georgiana Dunca et al. / Energy Procedia 112 (2017) 100 – 107

system at the time t 39 , corresponding to the 39th test in table 1. We plot in figure 5 the variation of the pumped flow rate through the hydraulic system, issued from the above TCV-numerical model: the pumps operate according to the speed patterns from table 1, and the total flow rate values fit the values from table 1.

Fig. 3. TCV-numerical model of the pumping station test rig; on each pipe, the length is displayed in meters (from 0.1 m to 0.66 m).

Fig. 4. Flow rate distribution obtained using the TCV-numerical model, at time moment t

39 .

Fig. 5. Pumped flow rate for each pump and total pumped flow rate at the discharge valve (TCV), obtained with the TCV-numerical model.

According to the Energy Report delivered by EPANET, the pumps P1÷P3 were used (put in operation) during the experiments with the following percent utilization values: P1 by 77.78%, P2 by 60.32% and P3 by 57.14% (the

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Georgiana Dunca et al. / Energy Procedia 112 (2017) 100 – 107

above percents are expressed with respect to the total duration, 100% being attached to all 63 time moments). The total power consumed for pumping (computed outside EPANET) was equal to 12.90 kW for the pump P1, 9.32 kW for P2, 8.50 kW for P3, and 30.72 kW for all 3 pumps (meaning for the entire pumping station test rig). Due to the important energy consumption, additional results obtained through numerical simulations can be useful within the Hydraulic Machinery Laboratory, where the experimental test rig operates. To extend the results, we conceived further a second numerical model, able to simulate the pumping station operation for any choice of the speed patterns, using a discharge valve able to ensure the requested flow rate, based on a curve ] ] Q . After analysing the minor loss coefficient values from table 1, we found that ] varies upon the total pumped flow rate under the following regression curve: ] ] Q a Qb , where a # 10017 and b 2.042 for the flow rate in m3/h (the proposed regression curve applies only to the studied TCV). Starting from the above dependency ] ] Q , we built an alternative numerical model in EPANET, called further as GPV-numerical model, by replacing the Throttle Control Valve (TCV) with a General Purpose Valve (GPV). For that new model, all simple control statements were removed, and we set at the GPV the curve describing the head losses h loss versus flow rate [3], [14], issued from the following relationship:

h loss

f Q 0.0826

] Q § Q · D4

2

¨ ¸ # 24.94Q  0.042 , © 3600 ¹

(3)

where the flow rate is in m3/h, the head loss is in metres, the valve diameter is D 0.04 m and the minor loss coefficient depends on Q upon the previous proposed regression curve. Finally, through simulations performed with the GPV-numerical model, we obtained new duty points and consequently new parallel pumps operation points, then we plot them on a chart (figure 6), together with the experimental operation points from table 1. All those numerical and experimental points fit into the operation area attached to the applied pumping schedule.

Fig. 6. Operation area (coloured in light blue) attached to the above pumping schedule. Experimental and GPV-numerical duty points, together with the corresponding parallel pumps operation points.

Georgiana Dunca et al. / Energy Procedia 112 (2017) 100 – 107

The pumping operation area (figure 6) consists of the union of 3 curvilinear polygons, bordered upper and lower by the head limits H max 24.7 m and H min 18.1 m. The left side polygon fits between the performance curve of a pump running at minimal speed nmin 0.792n0 and the curve of a pump running at nominal speed n0 . The polygon in the middle fits between the pumping head  flow rate curves plotted for 2 pumps in parallel, namely the curve of one pump at nmin coupled with another at n0 , and the curve of parallel running at n0 . The right side polygon fits between the curves plotted for 3 pumps in parallel, namely the curve of one pump at nmin coupled with 2 pumps at n0 , and the curve of parallel running at n0 . As one can notice from figure 6, the new numerical results extend the number and distribution zones for the operating points, with respect to the experimental data. 4. Conclusions The present study relies on experimental tests conducted on a pumping station test rig found in a Hydraulic Machinery Laboratory at the University "Politehnica" of Bucharest. To extend the study, numerical simulations were performed in EPANET, using two numerical models: the first model (TCV-numerical model) was calibrated upon experimental data, while the second one (GPV-numerical model) was built in accordance with a head losses versus flow rate relationship, derived from the first numerical model. Due to the calibration, a perfect match resulted between experimental duty points and numerical points obtained with the first EPANET model. Additional duty points obtained with the GPV-numerical model fit into the operation areas attached to the pumping schedule, together with the experimental points. The proposed approach proved to be useful to study the pumps operation within the above hydraulic system, with a minimal experimental effort. Acknowledgements Thanks are due to Multigama Tech SRL  KSB Representative in Romania, for the booster system called "Hydro-Unit Utility Line" [12] studied in this paper, offered to the Hydraulic Machinery Laboratory from the Hydraulics, Hydraulic Machinery and Environmental Engineering Department, of the Power Engineering Faculty, University "Politehnica" of Bucharest. This work has been supported by the Executive Agency for Higher Education, Research, Development and Innovation, PN-II-PT-PCCA-2013-4, ECOTURB Project. References [1] Georgescu A-M, Georgescu S-C, Cosoiu CI, Hasegan L, Anton A, Bucur DM. EPANET simulation of control methods for centrifugal pumps operating under variable system demand. Procedia Eng 2015;119:1012-9. [2] Georgescu S-C, Georgescu A-M, Hydraulic networks analysis using GNU Octave (in Romanian). Bucharest: Printech Press; 2014. [3] Georgescu S-C, Georgescu A-M. EPANET Manual (in Romanian). Bucharest: Printech Press; 2014. [4] Horowitz FB, Lipták BG. Pump Controls and Optimization, in: Instrument Engineers’ Handbook: Process Control, 3rd edition, BG Lipták (ed.), Butterworth-Heinemann Ltd 1995:1362-1386. [5] Georgescu S-C, Georgescu A-M. Pumping station scheduling for water distribution networks in EPANET. UPB Sci. Bull, Series D 2015; 77(2):235-246. [6] Georgescu S-C, Georgescu A-M, Madularea RA, Piraianu V-F, Anton A, Dunca G. Numerical model of a medium-sized municipal water distribution system located in Romania. Procedia Eng 2015;119:660-8. [7] Georgescu A-M, Perju S, Georgescu S-C, Anton A. Numerical model of a district water distribution system in Bucharest. Procedia Eng 2014; 70:707-14. [8] Georgescu S-C, Popa R, Georgescu A-M, Pumping stations scheduling for a water supply system with multiple tanks. UPB Sci. Bull, Series D 2010;72(3):129-140. [9] Dunca G, Isbasoiu EC, Calinoiu C, Bucur DM, Ghergu C. Vibrations level analyse during pumping station Gâlceag operation. UPB Sci. Bull, Series D 2008;70(4):181-190. [10] Dunca G, Bucur DM, Isbasoiu EC, Calinoiu C. Transient behavior analysis. Study case: pumping station Gâlceag. UPB Sci. Bull, Series C 2007;69(4):651-8. [11] Georgescu A-M, Georgescu S-C, Petrovici T, Culcea M.. Pumping stations operating parameters upon a variable demand, determined numerically for the water distribution network of Oradea. UPB Sci. Bull, Series C 2007;69(4):643-50. [12] ***. Hydro-Unit Utility Line documentation: HU-2-DPVME6/4-B-DPC-DOL. DP-Pumps, The Netherlands; 2012. [13] Georgescu A-M, Cosoiu C-I, Perju S, Georgescu S-C, Hasegan L, Anton A. Estimation of the efficiency for variable speed pumps in EPANET compared with experimental data. Procedia Eng 2014;89:1404-1411. [14] Rossman L. EPANET 2 Users Manual. U.S. Environmental Protection Agency: EPA/600/R-00/057. Cincinnati, OH, USA; 2000.

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