Introduction

Centrifugal sewage pump failures can

have significant economic and environmen-

tal consequences due to the continuous flow

of sewage and the disruptions caused by

necessary repairs. These repairs often involve

dismantling the pipeline and pump, removing

blockages, and reassembly, all of which re-

quire specialized personnel and incur signifi-

cant costs. Therefore, acrucial design consid-

eration for sewage pumps is resistance to

blockages caused by materials present in the

sewage. While semi-open vortex impellers

offer some blockage resistance, they suffer

from low efficiency. Pre-shredders can miti-

gate blockages but are often expensive and

impractical for high-capacity pumps. In such

cases, reducing the number of impeller

blades (e.g., using two-blade or single-

blade designs) increases free passage,

a critical parameter for handling solids in

contaminated liquids such as sewage.

Free passage is defined as the maximum

diameter of a sphere that can pass unob-

structed through the pump’s hydraulic system.

However, the primary cause of impeller

clogging in sewage pumps is often not large

solids, but rather fibrous materials such as

towels, clothes, and diapers [1], which tend

to accumulate on the impeller inlet edge.

To mitigate this, many sewage pumps

incorporate a stationary front shroud with

cutting grove (those pumps are often called

a“contrablock”), which grinds fibrous mate-

rials and prevents their accumulation on the

impeller blades. While this significantly im-

proves blockage resistance, it can introduce

minor efficiency losses due to increased leak-

age flows between the stationary shroud and

the rotating impeller. Studies by Souza et al.

[2] and Caruso et al. [3] have investigated

P ompy, armatura/Pumps, fittings

Multiobjective optimization of efficiency, head,

free passage and vibrations of two blade impeller

for sewage pump with stationary front shroud

Wielokryterialna optymalizacja sprawności, wysokości podnoszenia,

swobodnego przelotu i poziomu wibracji pompy ściekowej

z wirnikiem dwułopatowym współpracującym z nieruchomą tarczą przednią

MIKOŁAJ SITNIEWSKI, JANUSZ SKRZYPACZ, PRZEMYSŁAW SZULC, MARCIN JANCZAK, WITOLD LORENZ

DOI 10.36119/15.2025.3.7

This paper presents a multiobjective optimization procedure for a two-blade impeller in a sewage pump with a

stationary front shroud. The optimization variables were the wrap angles on the front and back shroud, while the

meridional cross-section and volute remained constant. The objective was to maximize efficiency, head, and free

passage (e.g., minimizing blockage) while minimizing vibrations by reducing radial forces. A full factorial design with

five levels was used for CFD (computational fluid dynamics) calculations. After selecting the optimal design, the results

of initial and optimized impeller were validated experimentally. Numerical and experimental results demonstrated

good agreement. The optimized impeller showed improvements: head increased by 38,4%, efficiency by 9,5%, and

free passage by 8,3%. While CFD calculations predicted a reduction in averaged radial forces, vibration

measurements on volute indicated an increase for the optimized impeller by 46,1%. This discrepancy suggests that

radial forces may not be the primary factor influencing vibrations in this specific pump configuration. The findings of

this study contribute to the development of more efficient and reliable sewage pumps.

Keywords: sewage pump, optimization, two blade impeller, radial forces, vibrations

Artykuł przedstawia procedurę optymalizacji wielokryterialnej dwułopatkowego wirnika w pompie ściekowej z nie-

ruchomą tarczą przednią. Zmiennymi optymalizacyjnymi były kąty opasania na tarczy przedniej i tarczy tylnej, pod-

czas gdy przekrój meridionalny i spiralny korpus zbiorczy pozostały stałe. Celem było zmaksymalizowanie sprawno-

ści, wysokości podnoszenia i swobodnego przelotu przy jednoczesnej minimalizacji poziomu wibracji poprzez

redukcję sił promieniowych. Do obliczeń CFD (computational fluid dynamics) zastosowano pełny plan czynnikowy z

pięcioma poziomami. Po wybraniu optymalnego projektu, parametry pompy z wirnikiem początkowym i zoptymali-

zowanym zostały zweryfikowane eksperymentalnie. Wyniki numeryczne i eksperymentalne wykazały dobrą zgod-

ność. Zoptymalizowany wirnik wykazał poprawę: wysokość podnoszenia wzrosła o 38,4%, sprawność o 9,5% i

swobodny przelot o 8,3%. Podczas gdy obliczenia CFD przewidywały redukcję uśrednionych sił promieniowych,

pomiary poziomu wibracji na spirali wskazały na wzrost dla zoptymalizowanego wirnika o 46,1%. Ta rozbieżność

sugeruje, że siły promieniowe mogą nie być głównym czynnikiem wpływającym na poziom wibracji w tej konfigura-

cji pompy. Wyniki przedstawionych badań przyczyniają się do rozwoju pomp ściekowych o lepszych parametrach

energetycznych i użytkowych.

Słowa kluczowe: pompa ściekowa, optymalizacja, wirnik dwułopatowy, siły promieniowe, drgania

mgr inż. Mikołaj Sitniewski https://orcid.org/0009-0001-3081-8802, dr hab. inż. Janusz Skrzypacz https://orcid.org/0000-0003-2021-3487,

dr inż. Przemysław Szulc https://orcid.org/0000-0003-1753-9611 ‒ Politechnika Wrocławska

dr inż. Marcin Janczak, dr inż. Witold Lorenz https://orcid.org/0009-0005-5274-7127 ‒ Hydro Vacuum S.A.

Adres do korespondencji/Corresponding author: mikolaj.sitniewski@pwr.edu.pl

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Pompy, armatura

the impact of the gap between the impeller

and shroud on the performance of single-

blade impellers. These studies demonstrate

that a well-maintained gap minimizes effi-

ciency losses, emphasizing the importance of

regular gap adjustments throughout the

pump’s lifecycle. Yang Y. et al. [4] further in-

vestigated the effects of tip clearance in

three-blade impellers, describing the forma-

tion of vortices due to leakage flows. Their

research suggests that adjusting the clear-

ance can influence pump performance.

Overall, while the use of a stationary front

shroud may slightly decrease efficiency, its

benefits in preventing blockages make it

a valuable component in many sewage

pump designs.

Reducing the number of blades is benefi-

cial for increasing free passage, but it pres-

ents challenges during the design stage.

Standard design theories often assume uni-

form flow patterns, which are not present in

two-blade or single-blade impellers. Mini-

mizing the number of blades is associated

with several issues, including high levels of

fluid-induced vibrations that increase as the

blade count decreases. Furthermore, vibra-

tions induced by the fluid in one-blade and

two-blade impellers can overlap with vibra-

tions caused by mechanical faults, such as

unbalance, bearing misalignment, and

cocked bearings, at synchronous (1x) and

twice synchronous (2x) frequencies. Tan L. et

al. [5] investigated the influence of blade

wrap angle on the hydrodynamic radial

force of asingle-blade impeller. They found

that increasing the wrap angle improved

head and efficiency but narrowed the high-

efficiency operating range. The radial forces

were strongly dependent on the wrap angle,

with asignificant decrease observed as the

wrap angle increased. However, this study

primarily focused on energy characteristics

and did not include experimental validation

of the predicted radial forces or vibration

levels. Wang C. et al. [6] explored the impact

of rotation center eccentricity on radial force

in asingle-blade impeller, aiming to mitigate

large fluctuating forces. While they found that

eccentricity can effectively reduce radial

forces, their study also lacked experimental

validation of the predicted force reductions

and their impact on vibration levels. Kim et al.

[7, 8, 9, 10, 11] and Nguyen et al. [12] opti-

mized asingle-channel impeller to improve

efficiency and reduce unsteady radial forces,

thereby mitigating flow-induced vibrations.

Their optimization process involved modify-

ing the impeller and volute shapes using

Stepanoff’s theory and employing ahybrid

particle swarm optimization and genetic al-

gorithm coupled with surrogate modelling.

CFD simulations, including both steady and

unsteady Reynolds-Averaged Navier-Stokes

equations, were used to evaluate the de-

signs. The optimized design demonstrated

improved efficiency and reduced variable

radial forces, with experimental results con-

firming increased efficiency and decreased

vibration at the blade passing frequency

(BPF). Tan L. et al. [13] investigated asingle-

blade impeller using numerical simulations,

performance tests, and particle image velo-

cimetry (PIV). Their study focused on charac-

terizing secondary flows within the impeller,

including jet-wake structures at the impeller

outlet, which are known to contribute to

losses and induce vibrations. Cui et al. [14]

investigated the correlation between radial

forces, total entropy generation (TEG), en-

tropy generation rate (EGR), and flow-in-

duced vibrations in a single-stage, low-

speed centrifugal pump with a5-blade im-

peller across various operating conditions.

They employed CFD and fluid-structure inter-

action (FSI) analyses and validated their nu-

merical model through experimental mea-

surements of vibrations and energy parame-

ters. The results demonstrated that radial

force, TEG, EGR, and vibration levels were

minimized at the pump’s optimal efficiency

point and increased as operating conditions

deviated from this point. This finding supports

the use of radial force, TEG, and EGR as in-

dicators of flow-induced pump vibrations.

The study also confirmed the previously ob-

served phenomenon of changes in radial

force vector direction with variations in effi-

ciency. Pei et al. [15] utilized FSI simulations

with strong two-way coupling to investigate

unsteady flow-induced impeller oscillations

in a single-blade pump under off-design

conditions. To validate their CFD-FSI model,

they employed proximity sensors to measure

shaft and impeller deformations in the radial

direction and conducted experiments under

various operating conditions. While the strain

values obtained from FSI and experiments

showed good agreement, a phase shift of

half a shaft revolution was observed be-

tween the predicted and measured strain

values. Song et al. [16] investigated pressure

oscillations and radial forces in centrifugal

pumps with single and double-suction impel-

lers. Their results demonstrated good agree-

ment between CFD simulations and experi-

mental data in terms of Q-H characteristics

and pressure oscillations around the spiral

circumference. Furthermore, the study

showed that double-suction impellers effec-

tively reduce pressure oscillations and radial

forces. It is important to note that vibration

problems are not limited to single-blade im-

pellers. Two-blade impellers are also suscep-

tible to significant vibration levels, and exist-

ing research on this specific configuration is

relatively limited. Ma et al. [17] conducted

amulti-objective optimization of atwo-blade

sewage pump impeller, focusing on maxi-

mizing free passage, head, and efficiency.

The optimization variables included impeller

blade angle and volute cross-sectional area,

determined using Stepanoff’s theory. The

optimization process utilized CFD methods in

conjunction with amachine-learning-based

artificial neural network. While the study

achieved improvements in the target hydrau-

lic parameters, it did not explicitly consider

flow-induced vibrations, acritical factor for

two-blade impeller designs. Ren Y. et al. [18]

employed a genetic algorithm to optimize

atwo-blade sewage pump impeller, consid-

ering head, efficiency, and average wear

depth as optimization objectives. Their ap-

proach utilized a coupled Computational

Fluid Dynamics-Discrete Element Method

(CFD-DEM) simulation.

As demonstrated by these previous studies,

effective methodologies for designing two-

blade impellers that simultaneously achieve

increased free passage, high efficiency, and

reduced vibrations are currently lacking. This

research addresses this gap by presenting

a multi-objective optimization framework for

two-blade impellers based on logarithmic

curves (curves with aconstant beta angle).

Methods

A Hydro-Vacuum S.A. pump model

FZF.5.21, equipped with aDN 150 inlet and

aDN 125 outlet in accordance with the EN-

1092 standard, was used for the tests. The

impeller was modified, while the mechanical

drive, volute, and stationary front shroud re-

mained unchanged. Figure 1 illustrates the

Figure 1.

Meridional section and axial view with marked maximum free passage value

Rysunek 1. Przekrój merydionalny iwidok na płaszczyznę prostopadłą do osi zzaznaczoną mak-

symalną wartością swobodnego przelotu

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meridional and axial view of the impeller. The

geometric parameters are summarized in

Table 1.

Table 1. Main geometrical parameters of impeller

Tabela 1. Główne wielkości geometryczne wir-

nika

Main geometrical parameters: Value:

Inlet diameter of impeller d

150 mm

Outlet diameter of impeller d

375 mm

Outlet width of the impeller b

80 mm

Blade wrap angle on front shroud j

To be defined

Blade wrap angle on back shroud j

To be defined

Diameter of free passage d

To be optimized

The research conducted by [19] indicates

that logarithmic curves offer asuperior blade

profile for centrifugal pump impellers. There-

fore, this study adopts the logarithmic curve

as the basis for blade design. The equation

describing a blade with a constant angle

along the curve is:

r = ae

(1),

where r is radius, a, b are constants, j is

angle.

Literature review presented above and

preliminary tests indicate that the wrap angle

is acritical parameter for impeller design. This

analysis investigates the influence of wrap

angle, implemented using logarithmic spirals

with a constant curve angle. To prevent fi-

brous materials from accumulating on the

impeller’s inlet edge, the edge must be ap-

propriately shaped, inclined towards the

stationary front disc. Our previous prelimi-

nary research demonstrated that an edge

that does not effectively deflect solids to-

wards the grinding disc can lead to the ac-

cumulation of fibrous materials. Therefore, in

this analysis, the inlet edge geometry was

kept constant, and the logarithmic curve on

the front disc was shifted by 30 degrees to

achieve the desired inlet edge inclination.

Afull factorial design was employed to

plan the experiment, with five levels for front

shroud wrap angles: 205°, 220°, 235°,

250°, 265° and five levels for back shroud

wrap angle 190°, 205°, 220°, 235°,

250°. This resulted in 25 simulations (5 levels

x 5 levels) to be conducted at the optimal

operating point of 340 m

/h.

Figure 2 shows the boundary conditions

for CFD calculations. CFD calculations were

performed using the Reynolds-Averaged Na-

vier-Stokes equations with the SST k-ω turbu-

lence model. All simulations were executed

on acomputing cluster using Fluent 2024 R1

software. Initially, 6,000 iterations of steady-

state calculations were performed, requiring

approximately 12 hours. Subsequently, tran-

sient simulations were conducted over three

impeller revolutions with a time step of

0.0001149 seconds, corresponding to aro-

tation angle of 1 degree at arotational speed

of 1450 RPM. This stage required approxi-

mately 48 hours of computation time. The use

of acomputing cluster facilitated the efficient

evaluation of 25 different geometries without

compromising mesh quality. The gap be-

tween the disc and the impeller was neglect-

ed in the CFD simulations. Numerical calcula-

tions did not incorporate frictional losses asso-

ciated with bearings, mechanical seals, and

the rotating rear disc, as these components

were not explicitly modelled. Consequently,

amechanical loss factor of 0.88 was applied

as amultiplier to the pump efficiency η.

Agrid independence study was conduct-

ed to ensure that the CFD results were not influ-

enced by mesh resolution. The mesh depen-

dence was evaluated at the best efficiency

point (340 m

/h). Figure 3 a) illustrates the

variation of efficiency with the number of mesh

cells, while Figure 3 b) shows the correspond-

ing variation in head. Based on these results,

amesh with 3.1 million elements was selected

for the subsequent analyses. Cross sectional of

that mesh with inflation layer is presented on

Figure 4.

In addition to efficiency, head, and ra-

dial force, which were evaluated using CFD,

the free passage of each impeller was as-

sessed using CAD software. The meridional

cross-section limits the free passage to 78

mm, as illustrated in Figure 1.

Following an approach described in the

literature review [5, 6, 8, 9, 10, 11, 12], the

optimization strategy focused on minimizing

vibrations by minimizing the radial forces act-

ing on the impeller. The optimization objec-

tives included: head, efficiency, free pas-

sage, and radial force. The weighted criteria

method was employed for optimization,

where each objective was first standardized

according to Equation (2). Each objective

was assigned an equal weight of 0.25.

(2)

z – standardized variable

x – non-standardized variable

m – mean of the population

s – standard deviation of the population

Results

Figure 5-8 present the results for the

head, the efficiency, free passage and the

radial force as afunctions of the front and

back shroud wrap angles. The results indicate

that the head is most significantly influenced

by the equality of front and back shroud

wrap angles, with the specific values of these

angles having a lesser impact. Figure 6

shows that impeller efficiency generally in-

creases with increasing wrap angles on both

the front and back shroud. The free passage

values increase with decreasing wrap an-

gles, but this trend is constrained by the

maximum allowable value determined by the

Figure 2.

Boundary conditions and cell zones for CFD

calculations

Rysunek 2. Warunki brzegowe iobjętości brył

wobliczeniach CFD

Figure 4.

Cross sectional of volute mesh

Rysunek 4. Przekrój siatki obliczeniowej spiral-

nego kanału zbiorczego

Figure 3.

Mesh dependency test: a) efficiency, b) head

Rysunek 3. Test wpływu wielkości siatki na:

a) sprawność, b) wysokość podnoszenia

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Pompy, armatura

meridional cross-section. The radial force ex-

hibits atrend similar to head, but since mini-

mizing radial force is akey objective, these

trends create aconflicting design requirement.

Figure 9 illustrates the optimization func-

tion after standardizing the variables and

summing the weighted criteria. The optimiza-

tion function reaches amaximum at a front

wrap angle of 250 degrees and a back

wrap angle of 216 degrees. The optimized

impeller was then re-simulated using CFD to

compare its performance with the initial im-

peller designs. CFD results were further vali-

dated through experimental testing. Figure

10 presents acomparison of radial forces for

the initial and optimized impellers during one

revolution, demonstrating a30% reduction in

radial force for the optimized design.

To validate the CFD calculations, the

optimized impeller design was 3D printed

using FDM technology. The initial design was

Figure 6.

3D plot of efficiency as a function of wrap angles on front and back

shroud

Rysunek 6. Wykres 3D sprawności w funkcji kąta opasania na tarczy

przedniej itarczy tylnej

Figure 5.

3D plot of head as a function of wrap angles on front and back

shroud

Rysunek 5. Wykres 3D wysokości podnoszenia wfunkcji kąta opasania

na tarczy przedniej itarczy tylnej

Figure 8.

3D plot of maximum radial force as afunction of wrap angles on front and

back shroud

Rysunek 8. Wykres 3D maksymalnej siły promieniowej w funkcji kąta

opasania na tarczy przedniej itarczy tylnej

Figure 7.

3D plot of free passage value as afunction of wrap angles on front and

back shroud

Rysunek 7. Wykres 3D swobodnego przelotu wfunkcji kąta opasania na

tarczy przedniej itarczy tylnej

Figure 9.

3D plot of optimized variable as afunction of wrap angles on front and back shroud

Rysunek 9. Wykres 3D funkcji optymalizacji wzależności od kąta opasania na tarczy przedniej

itarczy tylnej

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manufactured as acast iron impeller. These

impellers were then integrated into astand-

ard FZ.5 pump produced by Hydro-Vacuum

S.A. with with a three-phase electric motor

with a power of 110 kW and a rotational

speed of 1480 RPM. Figure 11 presents pho-

tographs of the tested impellers. Figure 12

shows the tested pump with discharge and

suction pipelines on the test stand. The gap

between the impeller and the grinding disc in

the impellers was adjusted and was within

the range of 0.2-0.4 mm.

Figure 13 and Figure 14 illustrates the

Q-H and Q-η curves for the initial and opti-

mized impellers. The presented optimization

procedure effectively ensured that the maxi-

mum efficiency was achieved at the target

flow rate of 340 m

/h.

It is well-known that reducing the pump’s

specific speed generally leads to adecrease

in achievable maximum efficiency. However,

in this case, increasing the head, which conse-

quently reduces the specific speed, resulted in

an increase in efficiency.

The omission of leakage between the

impeller and the front shroud resulted in

amarginal shift of optimal efficiency toward

lower values for both the initial and opti-

mized impeller.

Figure 10.

Radial forces of initial and optimized impeller during one rotation at nominal capacity (340 m

/h)

Rysunek 10. Siły promieniowe wirnika początkowego i zoptymalizowanego podczas jednego

obrotu przy wydajności optymalnej (340 m

/h)

Figure 13.

Q-H curve of initial and optimized impeller – CFD calculations and experiment

Rysunek 13. Wykres wysokości podnoszenia wzależności od wydajności dla wirnika początkowe-

go ioptymalnego – obliczenia CFD ieksperyment

Figure 14.

Q-η curve of initial and optimized impeller – CFD calculations and experiment

Rysunek 14 Wykres sprawności wzależności od wydajności dla wirnika początkowego ioptymal-

nego – obliczenia CFD ieksperyment

Figure 11.

Impellers for experiments – left optimized, right

initial

Rysunek 11. Wirniki do badań – zlewej strony

zoptymalizowany, zprawej strony początkowy

Figure 12.

Tested pump with discharge and suction pipeli-

nes on the test rig

Rysunek 12. Badana pompa z rurociągąmi:

tłocznym issawnym

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Table 2 and 3 present acomparison of

numerically simulated and experimentally

measured head and efficiency values at three

flow rates: 260, 340, and 420 m

/h. The ta-

bles also include the relative error of the nu-

merical calculations. The maximum relative

error of 12.48% was observed in the impeller

efficiency at the highest flow rate of 420

/h. However, for the subsequent analysis,

the point of best efficiency (BEP) is of greater

significance. At the BEP, the relative error re-

mained below 5% in all cases. Discrepancies

between the computational fluid dynamics

(CFD) results and the experimental data are

attributable, in part, to the surface roughness

of the hydraulic channels within the spiral cas-

ing and the impellers. The lower roughness of

the FDM-manufactured impeller, coupled with

the alignment of layer lines with the primary

flow direction within most of the channel, con-

tributed to the observed reduction in error for

this specific impeller.

Vibration measurements were conducted

on the circumference of the casing body for

both the initial and optimized impellers at three

flow rates: 260 m

/h, 340 m

/h, and 420

/h. The results, presented in Figure 15,

show root-mean-square (RMS) vibration lev-

els. For all measurements, the dominant fre-

quency was the blade passing frequency

(BPF), which in this case is 50 Hz due to the

two-bladed impeller.

An unexpected observation was an in-

crease in vibration levels on the circumference

of pump discharge casing for the optimized

impeller despite the lower radial forces pre-

dicted by CFD calculations. Furthermore, both

the initial and optimized impellers exhibited

the lowest vibration levels at 420 m

/h. This

contradicts the general understanding of cen-

trifugal pump behavior, which typically ob-

serves minimum vibrations at the point of opti-

mal efficiency. The lowest vibration levels were

observed in the plane of the discharge and

suction pipelines, potentially due to the greater

stiffness of the system in that direction com-

pared to the perpendicular direction.

Summary

This paper presents amulti-objective opti-

mization method for two-blade sewage pump

impellers, considering free passage, maximiz-

ing efficiency, and minimizing vibrations ex-

pressed in terms of radial forces. Stationary

and transient CFD calculations were per-

formed for afull grid sampling of two param-

eters: front and back wrap angle, each with

five levels, to determine the optimal combina-

tion of these parameters. The performance of

the initial and optimized impellers was vali-

dated through energy tests (Q-H and Q-η

curves) and vibration measurements.

The optimization process resulted in an

improvement in efficiency, head and free

passage. However, the expected reduction

in vibrations was not observed. This discrep-

ancy could be attributed to errors in the CFD

simulations, inaccuracies in the vibration

measurements, or amore complex relation-

ship between radial forces and actual vibra-

tion levels than initially anticipated.

Despite the challenges in minimizing vi-

brations, this research provides afoundation

for anovel approach to designing sewage

pump impellers with enhanced free passage

and improved hydraulic performance. Future

research should include experimental mea-

surements of radial forces acting on the im-

peller designs and further validation of the

CFD results. If the CFD results are accurate,

identifying and refining a suitable vibration

indicator within the CFD simulations will be

crucial for predicting vibration intensity more

effectively.

Created using resources provided by

Wroclaw Centre for Networking and Super-

computing (http://wcss.pl)

LITERATURE

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H, Kawahara T, Miyagawa K, Saito S. An

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Science. 2014;22. doi:10.1088/1755-1315/

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[2] Souza B, Daly J, Niven A, Frawley P. Numeri-

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[3] Caruso F, Meskell C. Effect of the axial gap on

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[4] Yang Y, Zhou L, Bai L, Xu H, Lv W, Shi W, Wang

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[5] Tan L, Yang Y, Shi W, Chen C, Xie Z.s. Influence

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[6] Wang C, Tan L, Shi W, Chenm C, Francis E.

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[7] Kim J-H, Cho B-M, Kim Y-S, Choi Y-S, Kim K-Y,

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Table 2. Experimental and CFD results of head for initial and optimized impeller

Tabela 2. Wyniki eksperymentalne iCFD wysokości podnoszenia dla wirnika początkowego iopty-

malnego

Table 3 Experimental and CFD results of efficiency for initial and optimized impeller

Tabela 3. Wyniki eksperymentalne iCFD sprawności dla wirnika początkowego ioptymalnego

Head [m]

Initial Optimized

H cfd H exp Relative error H cfd H exp Relative error

260 m^3/h 36,6 36,2 1,09% 45,5 45,5 0%

340 m^3/h (BEP) 31, 1 29,7 4,50% 42,1 41,1 2,43%

420 m^3/h 24,5 22,0 10,20% 38,3 35,8 6,53%

Efficiency

Initial Optimized

η cfd η exp Relative error η cfd η exp Relative error

260 m^3/h 6 7,2 68,4 -1,79% 71, 2 70,0 1,69%

340 m^3/h (BEP) 68,5 66,4 3,07% 72,5 72,6 -0,14%

420 m^3/h 65,7 57, 5 12,48% 73,4 69,6 5,18%

Figure 15.

Measurement of vibrations on the circumference

of the volute – V

rms

values [mm/s]: a) initial, b)

optimized

Rysunek 15. Wyniki pomiaru poziomu drgań na

obwodzie spiralnego korpusu zbiorczego –

wartości V

rms

[mm/s]: a) początkowy, b) opty-

malny

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doi:10.3390/pr7100747

[9] Kim J-H, Choi Y-S. State-of-the-Art Design

Technique of a Single-Channel Pump for

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[10] Kim J-H, Ma S-B, Choi Y-S, Kim K-Y. Simultane-

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n Poz. 188 OBWIESZCZENIE MARSZAŁKA SEJMU RZECZYPO-

SPOLITEJ POLSKIEJ zdnia 5 lutego 2025 r. wsprawie ogłoszenia jed-

nolitego tekstu ustawy oochronie przeciwpożarowej (Ogłoszone dnia

13 lutego 2025 r.)

Ustawa obejmuje:

– Rozdział 1 Przepisy ogólne

– Rozdział 2 Zapobieganie pożarowi, klęsce żywiołowej lub inne-

mu miejscowemu zagrożeniu

Art. 3. 1. Osoba fizyczna, osoba prawna, organizacja lub instytucja

korzystające ze środowiska, budynku, obiektu lub terenu są obowiązane

zabezpieczyć je przed zagrożeniem pożarowym lub innym miejscowym

zagrożeniem.

– Rozdział 2a Rzeczoznawcy do spraw zabezpieczeń przeciwpo-

żarowych

Art. 11b. 1. Rzeczoznawcą może być osoba, która posiada:

1) kwalifikacje wymagane do wykonywania zawodu inżynier po-

żarnictwa lub posiada tytuł zawodowy inżynier lub magister inżynier

oraz

2) przygotowanie zawodowe potwierdzone egzaminem złożonym

zwynikiem pozytywnym.

– Rozdział 3 Organizacja ochrony przeciwpożarowej

Art. 13. 1. Minister właściwy do spraw wewnętrznych określi, wdro-

dze rozporządzenia, sposoby i warunki ochrony przeciwpożarowej

budynków, innych obiektów budowlanych iterenów.

– Rozdział 4 Działanie ratownicze

– Rozdział 5 Uprawnienia strażaków jednostek ochrony przeciwpo-

żarowej

– Rozdział 6 Świadczenia rzeczowe ifinansowe

Art. 30. Właściciel, zarządca lub użytkownik budynku, obiektu lub

terenu ponosi wpełni koszty nabycia iutrzymania, wstanie zapewnia-

jącym sprawność, sprzętu, urządzeń przeciwpożarowych, środków ga-

śniczych, urządzeń sygnalizacyjno–alarmowych iinnych urządzeń oraz

instalacji ochrony przeciwpożarowej, do których posiadania zobowią-

zują go przepisy wydane na podstawie art. 13 ust. 1 i3, atakże koszty

wykonania obowiązku określonego wart. 5.

– Rozdział 7 Przepisy przejściowe ikońcowe

Art. 44. Właściciel, zarządca lub użytkownik budynku, obiektu lub

terenu, wktórym funkcjonuje resortowa lub zakładowa straż pożarna albo

inna jednostka ochrony przeciwpożarowej, wtym ochotnicza straż pożar-

na, ponosi wszelkie koszty związane zutrzymaniem tych jednostek.

Art. 47. Ustawa wchodzi wżycie wciągu 14 dni od dnia ogłoszenia.

n Poz. 235 ROZPORZĄDZENIE MINISTRA SPRAW WEWNĘTRZ-

NYCH IADMINISTRACJI zdnia 21 lutego 2025 r. wsprawie kryteriów

uznawania obiektów budowlanych albo ich części za budowle ochron-

ne (Ogłoszone dnia 26 lutego 2025 r.)

Określono między innymi wymagania dla budowli ochronnych za-

bezpieczających przed:

1) skutkami klęsk żywiołowych wywołanymi przez silne wiatry, wtym

wichury, orkany itrąby powietrzne,

2) odłamkami amunicji, w tym bomb, pocisków i granatów, oraz

przed ostrzałem zbroni małokalibrowej,

3) obciążeniami spowodowanymi zagruzowaniem oraz spadający-

mi elementami konstrukcji iwyposażenia obiektu budowlanego,

4) promieniowaniem przenikliwym gamma zopadu promieniotwór-

czego,

5) długotrwałym oddziaływaniem zewnętrznym pożaru na budowlę

ochronną,

6) skutkami fali uderzeniowej wybuchu,

7) skażeniem środowiska wewnętrznego wbudowli na skutek działa-

nia środków chemicznych, biologicznych lub promieniotwórczych,

8) wstrząsem oddziałującym na konstrukcję oraz wyposażenie bu-

dowli ochronnej

– wzakresie przewidzianym dla danej kategorii odporności budowli

ochronnej.

n Poz. 264 OBWIESZCZENIE MARSZAŁKA SEJMU RZECZYPO-

SPOLITEJ POLSKIEJ z dnia 21 lutego 2025 r. w sprawie ogłoszenia

jednolitego tekstu ustawy oprzygotowaniu irealizacji inwestycji wza-

kresie elektrowni szczytowo-pompowych oraz inwestycji towarzyszą-

cych (Ogłoszone dnia 4 marca 2025 r.)

n Poz. 280 ROZPORZĄDZENIE MINISTRA KLIMATU IŚRODOWI-

SKA zdnia 5 marca 2025 r. wsprawie szczegółowych warunków udzie-

lania przez Narodowy Fundusz Ochrony Środowiska iGospodarki Wod-

nej pomocy publicznej na inwestycje wmagazynowanie energii elektrycz-

nej izwiązaną znimi infrastrukturę (Ogłoszone dnia 7 marca 2025 r.)

Pomoc jest udzielana wformie: 1) dotacji bezpośrednich, 2) pożyczek.

Pomoc jest udzielana pod warunkiem, że inwestycja nie została

rozpoczęta przed dniem 9 marca 2023 r.

Pomoc jest udzielana do dnia 31 grudnia 2025 r.

Rozporządzenie wchodzi wżycie zdniem następującym po dniu

ogłoszenia.

Przegląd prawny wg Dziennika Ustaw

P rzegląd prawny/Rewiev of the law

Księga3_25.indb 60Księga3_25.indb 60 20.03.2025 11:39:2620.03.2025 11:39:26