Fullydeveloped internal turbulent flow
From ThermalFluidsPedia
Since the turbulent boundary layer grows much faster than the laminar boundary layer, the lengths of hydrodynamic and thermal entrances for turbulent internal flow are also much shorter than those for laminar flow. When the Prandtl number of the fluid is on the order of 1 (e.g., air or water), the lengths of the hydrodynamic and thermal entrances are about 10 times of the diameter of the tube, i.e.,

The internal turbulent flow becomes fully developed after x > L_{H} or L_{T}. Similar to the laminar internal flow, we have for fully developed flow, and the momentum equation becomes

where

is the apparent or total shear stress and y is the distance measured from the wall (y = r_{o} − r). The apparent shear stress is equal to τ_{w} at the wall and zero at the centerline. Integrating eq. (2) from the centerline to the wall yields,
which can be substituted into eq. (2) to yield

Integrating eq. (4) in the interval of (0, r), one obtains:

where y is measured from the tube wall (y = r_{0} − r). Equation (5) shows that the shear stress is a linear function for internal turbulent flow. Close to the wall where r is near r_{0} (or y is near 0), the apparent shear stress is nearly a constant, i.e., . The law of the wall resulting from the twolayer turbulent model (see Algebraic Models for Eddy Diffusivity) can be applied near the wall to yield:
u^{ + } = 2.5lny^{ + } + 5.5 
which is referred to as the Nikuradse equation. The constants 2.5 and 5.5 are different from those in Algebraic Models for Eddy Diffusivity and are obtained by curvefitting to the experimental results. The dimensionless velocity and coordinate are defined as

where

It should be pointed out that the Nikuradse equation (6) is invalid near the centerline because the slope of the velocity at the centerline obtained from eq. (6) is a finite value, not zero as it should be. In addition, eq. (6) also implies that at the centerline, which is also not true because the centerline is also in the fully turbulent region. Reichardt ^{[1]} suggested the following empirical correlation for the eddy diffusivity:

which becomes near the wall – a result that coincides with the mixing length theory. Equation (9) produces finite eddy diffusivity at the centerline. Assuming , eq. (3) becomes

Substituting eqs. (5) and (9), and integrating the resultant equation, the following velocity profile is obtained

which becomes identical to eq. (6) near the wall and produces zero slope at the centerline. The friction factor for internal turbulent flow is defined as

where is the mean velocity over the crosssection of the duct. For axisymmetric flow in a circular tube, it is obtained by

The definition of the friction factor, eq. (12), can be rewritten as

For moderate Reynolds number, the velocity profile in the entire tube can be approximated as ^{[2]}^{[3]}

At the centerline where , the centerline velocity satisfies:

The velocity at any radius is related to the centerline velocity by

Substituting eq. (17) into eq. (13), a relationship between the mean velocity and centerline velocity is obtained:

Substituting eq. (14) and (18) into eq. (16) and considering the definition of Reynolds number , the friction coefficient can be obtained as

which agreed with the experimental data very well up to . For even higher Reynolds number, the following empirical correlation works better for smooth tubes (see figure to the right):

Instead of the oneseventh law, eq. (15), the law of the wall, eq. (6), can be used to obtain the following correlation:

which is referred to in the literature as the KármánNikuradse relation. Equation (21) is valid up to Re_{D} = 10^{6}.
For fullydeveloped turbulent flow in a noncircular tube, eq. (21) is still applicable provided the hydraulic diameter is used in the definition of the Reynolds number. In this case, the friction coefficient, c_{f}, is defined based on the perimeteraveraged wall shear stress because the shear stress is no longer uniform around the periphery of the crosssection.
When the inner surface of the tube is not smooth, the friction will significantly increase with roughness (see figure to the right). The effect of the surface roughness can be measured by the roughness Reynolds number defined as:

where k_{s} is the roughness. When the roughness Reynolds number is greater than 70, the friction coefficient is no longer a strong function of the Reynolds number and becomes a constant, which is referred to as a fully rough surface. In the fully rough surface regime, the roughness size excees the order of the magnitude of what would have been the thickness of the viscous sublayer for a smooth surface. The friction coefficient for the fully rough surface regime can be obtained from the following empirical correlation:

References
 ↑ Reichardt, H., 1951, “Die Grundlagen des turbulenten Wärmeüberganges,” Arch. Gesamte Waermetech, Vol. 2, pp. 129142.
 ↑ Kays, W.M., Crawford, M.E., and Weigand, B., 2005, Convective Heat Transfer, 4th ed., McGrawHill, New York, NY
 ↑ Faghri, A., Zhang, Y., and Howell, J. R., 2010, Advanced Heat and Mass Transfer, Global Digital Press, Columbia, MO.