Jiří Škorpík, skorpik.jiri@email.cz

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1.22

Author:

ŠKORPÍK, Jiří, ORCID: 0000-0002-3034-1696

Issue date:

April 2010, May 2021, June, 2023 (3rd ed.)

Title:

Pressure loss at fluid flow and its calculation

Journal:

Transformační technologie (transformacni-technolgie.cz; fluid-dynamics.education; turbomachinery.education; engineering-sciences.education; stirling-engine.education)

ISSN:

1804-8293

Copyright©Jiří Škorpík, 2016-2023

All rights reserved.

Chapter: Description of pressure loss development and basic concepts

1.3

During the fluid flow, friction on the surface of the flow channel and the flowing bodies as well as friction inside the fluid (so-called internal friction). Through friction, the fluid loses kinetic energy and in order to flow through the channel at the required velocity (flow rate), it must gain kinetic energy at the expense of pressure energy - a pressure loss *L*_{p} is created, or at the expense of other energy, such as potential energy, etc.

Figure 1 shows the simplest case of pressure drop occurring when an incompressible fluid flows in a constant cross-sectional area pipe. Since the inlet and outlet of the duct must have the same flow, and therefore velocity, with no change in potential energy, the pressure drop *L*_{p} is equal to the pressure drop between the inlet and outlet, see Equation 1a.

The loss heat *L*_{q}, which is generated during friction, heats the working fluid. The loss heat for the case of Figure 1 corresponds to the pressure energy of the pressure loss, see Equation 1(b).

The fluid acts with a frictional force *F* on the channel in the direction of flow. The friction force for the case of Figure 1 corresponds to the product of the pressure difference between the inlet and outlet of the channel (pressure loss) and the flow area of the channel, see Equation 1(c).

For normal technical practice, addressing pressure loss in piping networks with valves is essential.Determining the pressure loss helps to calculate the work of the pump or fan - part of this work is consumed by the pressure loss. Heat loss calculations are crucial in cryogenics for transporting liquefied gases via pipelines, preventing property loss or evaporation. In the body, blood flow loss is compensated by the heart's activity, with output linked to pressure differentials (diastolic and systolic pressures).

Chapter: Description of pressure loss development and basic concepts

1.4

During fluid transport, incompressible fluid theories suffice for calculation pressure loss due to stable density. Yet, for gas transport in lengthy pipelines, density variations occur. Here, pressure loss calculations rely on mean gas density [Mikula et al., 1974, p. 71] or more precisely on friction-induced pressure loss equations in later paper chapters.

The procedure for calculating the pressure drop in the channel under investigation depends on whether the channel is in laminar or turbulent flow. This can be determined by the value of the Reynolds number for the case in question, for the calculation of which it is necessary to know the mean flow velocity of the fluid, the characteristic lenght of the channel (in the case of pipes this is the diameter) and the value of the kinematic viscosity. If the value of the Reynolds number is less than the value of the critical Reynolds number, then the flow is likely to be laminar, if the value of the Reynolds number is less than the value of the upper critical Reynolds number, then the flow is likely to be turbulent.

We also identify the pressure loss in channels designed for the transformation of pressure and kinetic energy of the fluid such as nozzles and diffusers or profile cascades, but in these cases the pressure loss is defined indirectly - the problem of losses in these channels is described in the articles Flow of gases and steam through nozzles^{4.}, Flow of gases and vapours through diffusers^{5.}, Aerodynamika profilových mříží [Škorpík, 2022a].

The relationship for calculating the pressure loss for the case of laminar steady flow as function of dynamic pressure can be derived from the Navier-Stokes equations. This equation is called the Darcy-Weisbach equation, which was developed by the French engineer Henry Darcy (1803-1858) for pipelines, see Equation 2. Later, on the basis of long term experiments and deduction, the German engineer Julius Weisbach (1806-1871) confirmed the validity of this relationship for transient and turbulent flows and even for losses in pipe fittings and valves.

Chapter: Equations for calculating pressure loss in pipe

1.5

From the Darcy-Weisbach equation it follows that the pressure loss is a certain fraction of the dynamic pressure, this fraction is given by the value of the loss coefficient. For channels of constant flow area, the loss coefficient can be calculated using the equations given in the chapter Calculation of pipe loss coefficient. For other types of channels, such as elbows, valves, etc., see chapter Pressure loss coefficient of local losses.

The loss coefficient for a constant cross-section pipe is computed using Equation 3. It is therefore a function of the length and diameter of the pipe (*d* is taken to be the characteristic length if the pipe is of non-circular cross-section [Mikula et al., 1974, p. 91]) and a quantity called the friction coefficient.

The equation for the pipe friction coefficient in laminar flow *λ*_{LF} can be easily derived from the Navier-Stokes equations, see Equation 4. The determination of the value of the friction coefficient at turbulent flow is based on the conclusion of measurements made by Johann Nikuradse on a series of glass pipes with artificial roughness using a sand film. Nikuradse measured the pressure drop of several pipes with different relative surface roughnesses for selected Reynolds numbers and from there calculated the values of the friction coefficient *λ* according to the Darcy-Weisbach equation (Equation 2). From these values he produced a chart of the dependence of the friction coefficient on the Reynolds number and confirmed the existence of four regions with different dependences of the friction coefficient on the Reynolds number, see Figure 4.

Chapter: Calculation of pipe loss coefficient

1.6

left-practical division of Nikuradse chart into basic areas; right-view of original Nikuradse chart [Nikuradse, 1933]. (A) the friction coefficient is linear function of only Reynolds numbers without influence of pipe roughness - laminar flow region; (B) transition region of flow from laminar to turbulent - both laminar and turbulent flow can occur; (C) turbulent flow region, in which friction coefficient is function of both Reynolds numbers and relative roughness of pipe; (D) turbulent region, in which friction coefficient is function of relative roughness of pipe – the higher relative roughness, the greater coeicient of friction. *C* [1] relative pipe roughness, see also Nomogram 8; *Re* [1] Reynolds number; *Re*_{C} [1] critical Reynolds number; *λ*_{LF} [1] friction coefficient for laminar flow, see Appendix 4 for derivation of equation; *λ*_{SP} [1] friction coefficient for turbulent flow in hydraulically smooth pipes (*C*→0) [Bauer et al., 1950, p.148] ; *λ*_{RP} [1] limit from which friction coefficient does not change with increasing Reynolds number, the so-called flow in hydraulically rough pipe [Bašta, 2003, p. 23]; *ε* [m] absolute roughness of inner walls of pipe (for values see, for example, [Mikula et al., 1974], Table 7).

To calculate the friction coefficient in the regions (C-D) in Figure 4 (C-D) regions, semi-empirical relationships approximate Nikuradse chart values, supplemented by other measurements. An overview of these equations is given, for example, in [Štefan, 2009]. A universal and accurate equation for engineering, by Cyril Colebrook (1910-1997) [Míka, 1977, p. 150] (Equation 5), is central. The American engineer Lewis Moody (1880-1954) then created the widely used Moody chart, derived from the Colebrook equation [Cihelka et al., 1975, p. 684], [Roček, 2002, p. 230].

In region (C), turbulent velocity profile develops. In region (D), the evolution is already complete and even with increasing Reynolds number, the ratio of the kinetic energy of the fluid in the boundary layer to the kinetic energy in the flow core does not change.

The values of the marginal Reynolds numbers *Re*_{RP}, i.e. the approximate boundary between regions (C) and (D), can be calculated by substituting the equation for *λ*_{RP} into the Colebrook equation. Selected values of the marginal Reynolds numbers calculated in this way are given in Tabulce 6.

C | 1·10^{-6} | 1·10^{-5} | 1·10^{-4} | 0,001 | 0,01 | 0,01 | 0,04 | 0,05 | ||

Re_{RP} | 2,62·10^{9} | 2,22·10^{8} | 1,82·10^{7} | 1,42·10^{6} | 2,28·10^{5} | 1,02·10^{5} | 1,95·10^{4} | 1,48·10^{4} |

Chapter: Calculation of pipe loss coefficient

1.7

ε | ε | ||||||

Drawn pipes (new) from: Copper, Brass, glass | 0,001...0,002 | Cast iron | 0,2..0,6 | ||||

Plastic and ruber | 0,0015...0,007 | Galvanized steel | 0,07...0,1 | ||||

Steel | 0,04...0,1 | Corroded steel pipes cleaned | 0,15...0,2 | ||||

Weld steel | 0,04...0,1 |

The Darcy-Weisbach equation implies that for minimum pressure loss it is advantageous to transport gas at higher pressures and densities than at low pressures and high velocities. Therefore, the pressures in transit pipelines are around 7 MPa and the gas pressure is reduced before the appliances (see Table 9), which are designed for lower pressures for safety reasons.

p | p | |||||

Transit pipeline | 7,5 | Medium pressure gas pipeline | 0,1...0,3 | |||

High pressure gas pipeline | 4 | Low pressure (household) | 0,002 |

For basic pipe route designs, designers use the quantity specific pressure loss in the pipe corresponding to the pressure loss in a 1 m long pipe, see alsoNomogram 10.

Chapter: Pressure loss per unit lengt of pipe

1.8

The pipe route (pipe network) is not usually straight and may consist of other pipe elements (branch pipes of various shapes, bends, constrictions), fittings, filters, meters and other flow parts, see Figure 11. These elements are local resistances and local pressure loss occurs in them.

Chapter: Pressure loss coefficient of local losses

1.9

In local resistances, the pressure losses are similar to those in a straight pipe. These pressure losses tend to be much more intense than in a straight section of pipe due to the fact that the flow through these sections also changes the shape of the flow channel, the direction of the flow and often the fluid throttling. Inlets and outlets of the pipe can also be considered as a special case of local resistance. At the edges, the flow is usually unsteady and influenced by the shape of the beginning or end of the pipe.

The local resistance pressure loss can also be calculated using Equation 2. The calculation of the pressure loss occurring in a given local resistance is based on the mean flow velocity before the element and the loss coefficient of the local resistance.

The loss coefficient *ζ* of some types of local resistances can be calculated [Maštovský, 1964, p. 85], but more often it is based on measurements of the local resistance for different Reynolds numbers. However, for some types of local resistances the influence of the Reynolds number is not significant and tabulated values can be used, especially for valves and pipe fittings, e.g. in [Cihelka et al., 1975, p. 672], [Miller et al., 1972, p. 252], [Řasa and Švercl, 2004, p. 737]. The corresponding loss factor is provided by the manufacturer of the local resistor in question. Loss coefficients for different types of pipe edges are given in [Ibler et al., 2002, p. 268].

Chapter: Pressure loss coefficient of local losses

1.10

In the case of valves, the manufacturer usually also directly supplies charts of the dependence of their pressure loss on the flow rate (depending on the type of flowing medium). If the nominal flow coefficient of a *K*_{VS} valve is known, the loss versus flow can be calculated from the definition of the flow coefficient, see Equation 12. The nominal flow coefficient is measured on the 2·*d* pipe section upstream of the valve and the 8·*d* pipe section downstream of the valve, so the loss coefficient calculated in this way includes this length of pipe - so the actual loss coefficient of the valve is lower by the loss coefficient corresponding to a 10·*d* smooth pipe. Approximate values of loss coefficients of some valves are given in [Roček, 2002, p. 231, 232]. However, there are other types of coefficients, usually derived from the pressure loss of the valve. It depends on the manufacturer what methodology he uses to compare valves. The relevant relationships are then given in the valve catalogue.

When selecting the most suitable closing valve, the allowable pressure loss *L*_{p} is first determined at the volume flow rate *Q* and the density of the flowing medium at the inlet *ρ*. The nominal flow coefficient *K*_{VS} is calculated. Next, the valve with the next higher *K*_{VS} is selected from the valve catalogue of the relevant manufacturer.

For approximate calculation of the local resistance pressure loss, a quantity called the equivalent pipe length can also be used. This quantity gives the length of smooth pipe (expressed as the number of diameters of smooth pipe) of the same diameter as the input diameter of the local resistance under investigation with the same pressure loss. Equivalent pipe lengths of some valves and pipe fittings are given in [Izard, 1961], [Fraas, 1989], and a selection is given in Table 13. The advantage is that in the calculation it is sufficient to add the individual equivalent lengths and calculate their total pressure loss as if they were the same length of hydraulically smooth pipe, see Problem 1.

l·d^{-1} | l·d^{-1} | |||||

GLOBE VALVES | ||||||

with no obstruction | 340 | Y-pattern with stem 60° from run of pipeline | 175 |

Chapter: Pressure loss coefficient of local losses

1.11

l·d^{-1} | l·d^{-1} | |||||

with guided in flow area (under seat) | 450 | Y-pattern with stem 45° from run of pipeline | 145 | |||

ANGLE VALVES | ||||||

with no obstruction | 145 | with guided in flow area (under seat) | 200 | |||

GATE VALVES | ||||||

convetional wedge | 13 | conduit pipeline | 3 | |||

pulp stock | 17 | |||||

CHECK VALVES | ||||||

conventional swing | 35 | in-line ball | 150 | |||

clearway swing | 50 | foot valves with strainer with poppet lift-type disc | 420 | |||

globe | 340 | foot valves with strainer with leather-hinged disc | 75 | |||

angle | 145 | butterfly valves | 20 | |||

COCKS | ||||||

rectangular plug port area equal to 100% of pipe area | 18 | three-way | 140 | |||

FITTINGS | ||||||

90° standard elbow | 30 | square corner elbow | 57 | |||

45° standard elbow | 16 | 180° close pattern return bend | 50 | |||

90° long radius elbow | 20 | standard tee with flow through run | 20 | |||

90° street elbow | 50 | standard tee with flow through branch | 60 | |||

45° street elbow | 26 | |||||

FLOW MEASUREMENT | ||||||

turbine flow meter | 150 | orifice plates | 200 | |||

piston meter | 400 |

The Darcy-Weisbach equation links higher mean flow velocity to increased pressure loss. This, in turn, impacts the cost of acquiring and operating machinery (e.g., pump, fan). Larger pipeline diameters, reducing mean flow velocity, raise costs for pipeline routes and fittings. Economic velocity, balancing pressure loss and acquisition expenses, is precisely calculated in [Krbek et al., 1999, p. 187]. Usual economic velocities, balancing costs, are derived from this compromise [Mikula et al., 1974, p. 141], as shown in Table 14. However, factors like layout considerations may influence velocities beyond economic reasons.

Chapter: Economic velocity in pipe

1.12

V‾ | V‾ | |||||

oil | 1...2 | steam superheated to 4 MPa | 20...40 | |||

water | 1...4 | steam superheated at high pressure | 30...60, 80 | |||

low pressure heating steam | 10...15 | exhaust steam (after expansion in machine) | 15...30 | |||

sat. steam up to 1 MPa | 15...20 | air (compressed) | 2...15 |

The pipe diameter *d* is calculated from the design economic velocity, density and required specific flow rate, see Nomogram 15. The calculated pipe diameter must be rounded off according to the manufactured pipe diameters corresponding to the pressure and temperature at which the pipe will be operated.

Chapter: Economic velocity in pipe

1.13

The characteristic pipeline is the dependence of the pressure loss of the pipeline route on the volumetric flow rate. From the equation for calculating the pressure loss it is clear that at *ρ*=const. the pressure loss will be a quadratic function with a parameter *C*_{S} called the pipeline system constant, see Equation 16.

The piping system constant *C*_{S} is usually considered as a constant for a given opening of individual valves, but since the friction coefficient *λ* is a function of the Reynolds number, *C*_{S} must also change with the flow rate. However, this change is not very large if we are interested in the pressure loss in the nominal flow region. For calculations in a larger range of flow rates, a correction can be applied by not multiplying the volumetric flow rate by 2, but by another exponent, see more in [Bašta, 2003, p. 25].

The pipeline system constant can be calculated according to Equation 16 from the individual pressure losses of the pipeline system for a known (nominal) flow rate (see Problem 1) or it can be calculated from the measured pressure loss at a particular volumetric flow rate, see Problem 2.

Chapter: Determination of characteristics pipeline from measurements

1.14

The characteristic pipeline or fuction *L*_{p}=f(*Q*) can be established by measuring different cases. This data can then be processed using computer software or determined by plotting on logarithmic paper, creating a straight line whose slope corresponds to the flow rate exponent, as seen in Problem 2 and the article Engineering mathematics [Škorpík, 2023].

A fouling can form in the pipe if the liquid is not clean. Fouling in a pipeline system causes a reduction in the flow area of the pipeline and therefore a change in the characteristics pipeline and an increase in the pressure loss. Figure 17 shows the change in pressure loss in a pipeline when there is a uniform fouling in the pipeline - approximately the same percentage increase in pressure loss will increase the pumping work. The relationship in this figure was developed by substituting the Darcy-Weisbach equation into the pressure loss ratio *L*_{p} after reducing the flow area and the pressure loss *L*_{p,n}. From here it can be seen the reduction in diameter per pressure loss increases with the fifth power. On the other hand, even when absolute roughness is maintained, the effect of the change in friction coefficient is several orders of magnitude smaller.

Chapter: Change in pressure loss due to pipe fouling

1.15

Fouling of the pipeline may be caused by chemical or biological action or by solid particles in the liquid. In the case of a chemical or electrochemical process, minerals precipitate and crystallise on the internal surfaces of the pipe. The biofouling on the pipe can be of plant or animal origin - usually some kind of algae or crustacean - and are highly dependent on water temperature, nutrient content of the water and, in the case of algae, light conditions. A typical sign of fouling by solid particles in the liquid is that it is not evenly distributed along the length of the pipe. The solid particles are deposited in areas of low flow velocity, at the lowest points of the pipeline route where the fluid flow is unable to displace them and upstream of constrictions.

Scale deposition on the pipe walls does not occur at velocities of approximately 1,5 to 2,5 m·s^{-1} [Vosmík, 2023]. However, at certain combinations of pH and temperature, this velocity may not be sufficient. Deposition of solid particles can be prevented from velocities as low as around 1,5 m·s^{-1}, but also depends on the orientation of the pipe and the size and mass of the individual particles, according to [Pugh et al., 2009]. Biofouling of pipes can be prevented at velocities above 2 m·s^{-1}.

These velocities are for water. For other liquids, the limiting velocity may vary because a certain tangential tension, which is a function of viscosity, is required to prevent fouling at lower velocities and vice versa. Details on fouling of pipes and exchangers are given in the referenced paper [Pugh et al., 2009].

Constant flow velocity during irregular pipeline operation can be maintained by creating loops on the exposed parts of the pipeline in which the fluid will flow at a constant velocity regardless of the flow rate between the inlet and outlet of the pipeline. Alternatively, a significantly higher nominal flow rate must be provided at switch-on to clear the pipe after partial operation (e.g. after a night of light operation). Prevention of pipe fouling can also be done by modifying the working fluid, filters or changing temperatures, but this is beyond the scope of this article.

Chapter: Change in pressure loss due to pipe corrosion

1.16

Pipe corrosion increases the absolute roughness of the pipe and causes a loss of pipe wall thickness. If the material loss does not cause a significant change in the flow surface area of the pipe, then, given the other parameters in the Darcy-Weisbach equation, the ratio of the pressure loss *L*_{p} to the pressure loss at nominal (initial) *L*_{p,n} can be expressed as a ratio of the coefficients of friction. Figure 18 is a plot of the change in pressure loss in a steel pipe as roughness increases due to corrosion of the cleaned steel pipe (data from Table 7), showing that corrosion can increase the pressure loss by tens of percent.

In addition to fluid transport, we encounter dynamic gas flow in which the density of the gas can change significantly. If it is an adiabatic flow, then the pressure loss can be determined by assuming that the stagnation enthalpy of the gas remains constant and equal to the stagnation enthalpy at the inlet, but the entropy will increase due to internal friction. Based on this assumption, general Equations 19, which describes gas flow in the presence of friction in all types of channels, can be derived from the continuity, energy balance and momentum conservation equations for the assumption of constant specific heat capacity of the gas. However, in engineering practice, these equations are only used in calculations of flows with large density variations in narrow channel seals.

Chapter: Pressure loss at significant density change

1.17

The friction coefficient *λ* in Equation 19 is a constant along the length of the channel, but in actual fact is more or less dependent on *Re* and the Mach number at the channel location under investigation. Thus, it depends on how much the channel flow area and Mach number vary. Experimental verification of the changes in the friction coefficient during compressible flow and the validity of Equation 19 is carried out in [Dejč, 1967, p. 217].

In the case of compressible adiabatic flow in a channel of constant flow area, the pressure loss can be calculated using Equation 20, which is based on a modification of the general Equation 19 for the condition d*A*=0.

Chapter: Pressure loss at significant density change

1.18

In adiabatic gas flow, friction heats the gas, increasing its specific volume and velocity in a constant flow area channel. This leads to a gradual decrease in gas pressure and specific enthalpy. The Fanno line on the *h*-*s* diagram plots gas states along the channel axis. Figure 21 depicts three Fanno lines for a channel of length *l* with varying friction coefficients *λ*, influencing pressure changes as the channel lengthens (the same effect as changes in the friction coefficient has on the pressure change as the channel lengthens). At the maximum *λ*_{1}, outlet flow doesn't reach critical velocity; *λ*_{2} just reaches critical velocity, and *λ*_{3}, less than *λ*_{2}, also reaches critical velocity at the outlet.

In engineering practice, the theory is particularly applicable to the investigation of flow in non-contact seals. The principle of dry-running gas seals is also based on the high pressure loss associated with gas flow in a very small gap. However, even labyrinth seals can be likened to a smooth seal with a constant flow area and a particular coefficient of friction.

Chapter: Pressure loss at significant density change

1.19

Find the characteristics pipeline at the discharge of a condensate pump (see attached figure) in which condensate is pumped from the auxiliary condensate tank CT1 to the feed tank through the condensate heater H1. A parallel pipeline system with a redundant pump (blue) is connected to the route. The water temperature at the outlet of the pump is 60 °C and 105 °C after the H1 heater. The flow rate through the pump is 2,4 m^{3}·h^{-1}. The flow coefficient of ball valve 001 is 48,5 m^{3}·h^{-1}. The check valve has a pressure loss of 5 kPa. The minimum pressure loss of the balancing valve is 750 Pa. The pressure loss of the water meter is 18 kPa. The pressure loss of heater H1 is 12 kPa. The piping is standard one-inch water main. The solution to the problem is shown in Appendix 1.

CT1-auxiliary condensate tank No. 1; H1-heater No. 1; WM1-water meter No. 1. Marking according to [Krbek et al., 1999, p. 178]. The lengths of the individual sections of the piping system are given in metres.

Find the approximate value of the constant of the heating piping system. Hot water flows through the pipe. There are the measured flows through the system and the corresponding pressure loss given in the table below. Measured values adapted from [Pleskot, 1947, p. 17]. The solution to the problem is shown in Appendix 2.

Table of measured values

L_{p} | 10 | 25,1 | 62 | 140 | 320 | 700 | 1400 | |

Q | 19,64 | 29,64 | 50,07 | 74,61 | 113,9 | 161 | 233,7 |

Chapter: Problems

1.20

ŠKORPÍK, Jiří, 2022a, Aerodynamika profilových mříží, *Transformační technologie*, Brno, [online], ISSN 1804-8293. https://turbomachinery.education/aerodynamika-profilovych-mrizi.html.

ŠKORPÍK, Jiří, 2022b, Základní rovnice lopatkových strojů, *Transformační technologie*, Brno, [online], ISSN 1804-8293. https://turbomachinery.education/zakladni-rovnice-lopatkovych-stroju.html.

ŠKORPÍK, Jiří, 2022c, Úvod do lopatkových strojů, *Transformační technologie*, Brno, [online], ISSN 1804-8293. https://turbomachinery.education/uvod-do-lopatkovych-stroju.html#energeticka-bilance-lopatkoveho-stroje.

ŠKORPÍK, Jiří, 2023, Technická matematika, *Transformační technologie*, Brno, [online], ISSN 1804-8293. Dostupné z https://engineering-sciences.education/technicka-matematika.html.

BAŠTA, Jiří, 2003, *Hydraulika otopných soustav*, Vydavatelství ČVUT, Praha, ISBN 80-01-02808-9.

BAUER, František, Oldřich BRŮHA a Zbyněk JAŇOUR, PEŠEK, Rudolf, ed., 1950, *Základy proudění*, Vědecko-technické nakladatelství, Praha.

CIHELKA, Jaromír, BRANDA, Jaroslav, CIKHART, Jiří, ČERMÁK, Jan, CHYSKÝ, Jaroslav, PITTER, Jaroslav, VALÁŠEK, Jiří, 1975, *Vytápění a větrání*, SNTL, Praha.

DEJČ, Michail, 1967, *Technická dynamika plynů*, SNTL, Praha.

FRAAS, Arthur, 1989, *Heat exchanger design*, John Wiley&Sons, Inc., ISBN 0-471-62868-9.

IBLER, Zbyněk, KARTÁK, Jan, MERTLOVÁ, Jiřina, IBLER, Zbyněk ml., 2002, *Technický průvodce energetika*, BEN-technická literatura, Praha, ISBN 80-7300-026-1.

IZARD, Julien, 1961, *Příručka technické fyziky*, Státní nakladatelství technické literatury, Praha.

JÍCHA, Miroslav, 2001, *Přenos tepla a látky*, Vysoké učení technické v Brně, Brno, ISBN 80-214-2029-4.

KRBEK, Jaroslav, POLESNÝ, Bohumil, FIEDLER, Jan, 1999, *Strojní zařízení tepelných centrál-Návrh a výpočet*, PC-DIR Real, s.r.o., Brno, ISBN 80-214-1334-4.

MAŠTOVSKÝ, Otakar, 1964, *Hydromechanika*, Statní nakladatelství technické literatury, Praha.

MÍKA, Vladimír, 1977, *Základy chemického inženýrství*, Statní nakladatelství technické literatury, Praha.

MIKULA, Julius, KOČKA, Jaroslav, ŠKRAMLÍK, Emanuel, ŠTAUBER, Zdeněk, VESELÝ Adolf, OBR, Jan, 1974, *Potrubí a armatury*, Státní nakladatelství technické literatury, Praha.

Chapter: References

1.21

MILLER, Rudolf, HOCHRAINER, A., LÖHNER, K., PETERMANN, H., 1972, *Energietechnik und Kraftmaschinen*, Rowohlt taschenbuch verlag GmbH, Hamburg, ISBN 3-499-19042-7.

NIKURADSE, Johann, 1933, Strömungsgesetze in rauen Rohren, *V. D. I. Forschungsheft*, 361: 1–22, Berlin.

PLESKOT, Václav, 1947, *Nomografie v technické praxi*, Praha, SPASEI.

POLESNÝ, Bohumil a kol., 1990, *Termodynamická data pro výpočet tepelných a jaderných energetických zařízení*, Vysoké učení technické v Brně, Brno, ISBN 80-214-0160-5.

PUGH, Simon, HEWITT, Geoffrey, MÜLLER-STEINHAGEN, Hans, 2009, Fouling During the Use of “Fresh” Water as Coolant—The Development of a “User Guide”, *Heat Transfer Engineering*, 30:10-11, 851-858, DOI: 10.1080/01457630902753706.

ROČEK, Jaroslav, 2002, *Průmyslové armatury*, INFORMATORIUM, Praha, ISBN 80-7333-000-8.

ŘASA, Jaroslav, ŠVERCL, Josef, 2004, *Strojnické tabulky, jednotky, matematika, mechanika, technické kreslení, strojní součásti*, Scientia, spol. s.r.o. Praha, ISBN 80-7183-312-6.

ŠTEFAN, David, 2009, *Hydraulické ztráty v potrubí*, Vysoké učení technické v Brně, Brno, Bakalářská práce.

VOSMÍK, Zdenek, 2023, *Nerezové výměníky tepla – koroze a zanášení úsadami 2. část*, https://www.vosmik-vymeniky.cz/, https://www.vosmik-vymeniky.cz/novinky.html.

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