# Steady-state heat conduction in plane and radial walls

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## Generalized 1-D heat conduction

One-dimensional heat conduction can occur in different geometric configurations. The figure on the right shows heat conduction in a plane wall and in a hollow cylinder or sphere. The energy equationin different geometric configurations can be expressed as:

Plane wall:

Hollow cylinder:

Hollow sphere:

Equations (1) – (3) can be represented using the following general form (Arpaci, 1966):

where *s* is space variable and *A*(*s*) can be viewed as the heat transfer area. The heat transfer area in a plane wall, hollow cylinder and sphere are *A*(*s*) = *L**W*, 2π*s**L*, and 2π*s*2, respectively (where *W* is the width of the plane wall, and *L* is length of the plane wall or cylinder). Figure 2 represents the general one-dimensional heat conduction problem.

Equation (4) can also be rewritten as

or

which means the heat transfer rate at different cross-sections is a constant. It should be pointed out that the heat flux *q*''(*s*) = − *k**d**T* / *d**s* is not a constant unless the cross-sectional area *A*(*s*) is constant.
If the thermal conductivity is independent from temperature, the temperature distribution can be obtained by integrating eq. (4) twice:

where *a* and *b* are unspecified integral constants. If the temperatures of both inner and outer surfaces are specified, i.e.,

The integral constants *a* and *b* in eq. (7) can be determined from eqs. (8) and (9), and the temperature becomes

## Thermal Resistance

The rate of heat transfer can be obtained using Fourier’s law:

which can be rewritten as

where

is the thermal resistance for heat conduction. Equation (12) is similar to Ohm’s law

where *V*_{1} and *V*_{2} are, respectively, the electrical potential at two ends of the electrical resistance, *R*, and *I* is the current through the electrical resistance (see Fig. 3).

If the boundary conditions at the inner and outer surfaces are convective conditions, i.e.,

where *h*_{i} and *h*_{o} are heat transfer coefficients at the inner and outer surfaces, respectively.
The temperatures of the fluids that are in contact with the inner and outer surfaces are *T*_{i} and *T*_{o}, respectively.
If the temperatures of the inner and outer walls are represented by *T*_{1} and *T*_{2}, (both of them are unknown) eq. (12) is still valid. By multiplying eqs. (15) and (16) by the heat transfer areas at the inner and outer surfaces, *A*_{1} and *A*_{2}, one obtains

Under steady state condition, the heat transfer rate obtained from eqs. (12), (17) and (18) are identical. By eliminating *T*_{1} and *T*_{2} from these equations, the following expression for heat transfer rate is obtained:

where

are convective thermal resistances between inner fluid and inner wall and between outer fluid and outer wall, respectively. Equation (19) can be viewed as a system with three thermal resistances connected in series. Equation (19) can also be written as

where *U*_{o} is the overall- coefficient of heat transfer based on the outer surface area. Comparing eqs. (19) and (22) yields

or

For the heat conduction in three different coordinates as shown in eqs. (1) – (3), the overall coefficients for heat transfer are

If the conducting wall shown in Fig. 1 has multiple layers and each layer has different thermal conductivity, there will be multiple conduction thermal resistances between two convection thermal resistances. If the number of layers is represented by N, the overall coefficient of heat transfer will be expressed as

where *A*_{1} and *A*_{N + 1} are the areas of the inner and outer surfaces, respectively; and *k*_{i} is the thermal conductivity of the *i*^{th} layer (*s*_{i} < *s* < *s*_{i + 1}). The heat transfer rate becomes

or

## Contact thermal resistance

In arriving at eq. (28), it is assumed that different layers are in perfect contact so that the temperatures across the interfaces between different layers are continuous. When two rough surfaces are in contact, there may be a temperature drop, ΔT, across the interface between different materials (see Fig. 4). The heat flux and this temperature drop is related by

where *R*''_{ct} is contact thermal resistance (*m*2 − *K* / *W*). Heat transfer across an imperfect contact interface is due to combined effects of conduction at the actual contact area, convection of the air entrapped in the gap, and radiation between the two surfaces that are not in direct contact. The contact thermal resistance can be reduced by applying pressure in the direction perpendicular to the interface, or applying conduction grease at the interface. More information about contact thermal resistance can be found in Fletcher (1988).
When contact thermal resistances are present, eq. (30) can be modified as

## With Internal Heat Generation

The 1-D steady-state heat conductions that we discussed so far are limited to the case without internal heat generation. Heat conduction with internal heat generation can be encountered in many applications such as electrical heating, chemical reaction, or nuclear reaction in the conduction medium.
Let us consider the generalized 1-D heat conduction problem shown in Fig. 2. With uniform internal heat source, *q*''', the energy equation is

where the thermal conductivity is assumed to be independent from temperature. Equation (33) is subject to boundary conditions specified by eqs. (8) and (9). Multiplying eq. (19) by *A*(*s*) and integrating the resultant equation in the interval of (*s*_{1}, *s*) yields

Dividing eq. (34) by *A*(*s*) and integrating the resultant equation in the interval of (*s*_{1}, *s*), we have

where the temperature gradient at *s* = *s*_{1} is still unknown. Equation (35) already satisfies eq. (8) but not eq. (9). Substituting eq. (35) into eq. (9), one obtains

which can be rearranged to get

Therefore, the temperature profile becomes

The heat transfer rate can be obtained by Fourier’s law

Substituting eq. (34) into eq. (37), the heat transfer rate becomes

where

is the heat transfer rate at *s* = *s*_{1}. It is evident from eq. (38) that the heat transfer rate is no longer independent from s when there is internal heat generation.
It should be pointed out that eqs. (36) and (38) are valid for any coordinate system. For a Cartesian coordinate system with the origin of the coordinate on the left surface (i.e., *x*_{1} = 0 in Fig. 1), *A*(*s*) = *A* is a constant and eqs. (36) and (38) become

For the heat conduction in a cylindrical and spherical coordinate system, the general solution, eqs. (36) and (38), can be simplified by considering the variation of conduction area (see Problem 3.7 and 3.8).

## References

Arpaci, V.S., 1966, Conduction Heat Transfer, Addison-Wesley Pub. Co., Reading, MA.

Faghri, A., Zhang, Y., and Howell, J. R., 2010, *Advanced Heat and Mass Transfer*, Global Digital Press, Columbia, MO.

Fletcher, L.S., 1988, “Recent Developments in Contact Heat Transfer,” ASME J. *Heat Transfer*, Vol. 110, pp. 1059-1070.