Unit Operations in Food Processing

The basis of mechanical refrigeration is the fact that at different pressures the saturation (or the condensing) temperatures of vapours are different as clearly shown on the appropriate vapour-pressure/temperature curve. As the pressure increases condensing temperatures also increase. This fact is applied in a cyclic process, which is illustrated in Fig. 6.8.

The process can be followed on the pressure-enthalpy (P/H) chart shown on Fig. 6.9.

This is a thermodynamic diagram that looks very complicated at first sight but which in fact can make calculations straightforward and simple. For the present purposes the most convenient such chart is the one shown with pressure as the vertical axis (for convenience on a logarithmic scale) and enthalpy on the horizontal axis. On such a diagram, the properties of the particular refrigerant can be plotted, including the interphase equilibrium lines such as the saturated vapour line, which are important as refrigeration depends on evaporation and condensation. Figure 6.9 is a skeleton diagram and Appendix 11 gives charts for two common refrigerants, refrigerant 134a, tetrafluoroethane (in Appendix 11a), and ammonia (in 11b); others for common refrigerants can be found in the ASHRAE Guide and Data Books.

To start with the evaporator; in this the pressure above the refrigerant is low enough so that evaporation of the refrigerant liquid to a gas occurs at some suitable low temperature determined by the requirements of the product. This temperature might be, for example -18°C in which case the corresponding pressures would be for ammonia 229 kPa absolute and for tetrafluoroethane (also known as refrigerant 134a) 144 kPa. Evaporation then occurs and this extracts the latent heat of vaporization for the refrigerant from the surroundings of the evaporator and it does this at the appropriate low temperature. This process of heat extraction at low temperature represents the useful part of the refrigerator. On the pressure/enthalpy chart this is represented by ab at constant pressure (the evaporation pressure) in which 1 kg of refrigerant takes in (H_b - H_a) kJ . The low pressure necessary for the evaporation at the required temperature is maintained by the suction of the compressor.

The remainder of the process cycle is included merely so that the refrigerant may be returned to the evaporator to continue the cycle. First, the vapour is sucked into a compressor which is essentially a gas pump and which increases its pressure to exhaust it at the higher pressure to the condensers. This is represented by the line bc which follows an adiabatic compression line, a line of constant entropy (the reasons for this must be sought in a book on refrigeration) and work equivalent to (H_c - H_b) kJ kg^-1 has to be performed on the refrigerant to effect the compression. The higher pressure might be, for example, 1150 kPa (pressures are absolute pressures) for ammonia, or 772 kPa for refrigerant 134a, and it is determined by the temperature at which cooling water or air is available to cool the condensers.

To complete the cycle, the refrigerant must be condensed, giving up its latent heat of vaporization to some cooling medium. This is carried out in a condenser, which is a heat exchanger cooled generally by water or air. Condensation is shown on Fig. 6.9 along the horizontal line (at the constant condenser pressure), at first cd cooling the gas and then continuing along de until the refrigerant is completely condensed at point e. The total heat given out in this from refrigerant to condenser water is (H_c - H_e) = (H_c - H_a) kJ kg^-1. The condensing temperature, corresponding to the above high pressures, is about 30°C and so in this example cooling water at about 20°C, could be used, leaving sufficient temperature difference to accomplish the heat exchange in equipment of economic size.

This process of evaporation at a low pressure and corresponding low temperature, followed by compression, followed by condensation at around atmospheric temperature and corresponding high pressure, is the refrigeration cycle. The high-pressure liquid then passes through a nozzle from the condenser or high-pressure receiver vessel to the evaporator at low pressure, and so the cycle continues. Expansion through the expansion valve nozzle is at constant enthalpy and so it follows the vertical line ea with no enthalpy added to or subtracted from the refrigerant. This line at constant enthalpy from point e explains why the point a is where it is on the pressure line, corresponding to the evaporation (suction) pressure.

By adjusting the high and low pressures, the condensing and evaporating temperatures can be selected as required. The high pressure is determined: by the available cooling-water temperature, by the cost of this cooling water and by the cost of condensing equipment. The evaporating pressure is determined by either the low temperature that is required for the product or by the rate of cooling or freezing that has to be provided. Low evaporating temperatures mean higher power requirements for compression and greater volumes of low-pressure vapours to be handled therefore larger compressors, so that the compression is more expensive. It must also be remembered that, in actual operation, temperature differences must be provided to operate both the evaporator and the condenser. There must be lower pressures than those that correspond to the evaporating coil temperature in the compressor suction line, and higher pressures in the compressor discharge than those that correspond to the condenser temperature.

Overall, the energy side of the refrigeration cycle can therefore be summed up:

heat taken in from surroundings at the (low) evaporator temperature and pressure (H_b - H_a),

heat equivalent to the work done by the compressor (H_c - H_b) and

heat rejected at the (high) compressor pressure and temperature (H_c - H_e).

A useful measure is the ratio of the heat taken in at the evaporator (the useful refrigeration), (H_b - H_a) , to the energy put in by the compressor which must be paid for (H_c– H_b). This ratio is called the coefficient of performance (COP).
The unit commonly used to measure refrigerating effect is the ton of refrigeration = 3.52 kW. It arises from the quantity of energy to freeze 2000 lb of water in one day (2000 lb is called 1 short ton).

EXAMPLE 6.7. Freezing of fish
It is wished to freeze 15 tonnes of fish per day from an initial temperature of 10°C to a final temperature of -8°C using a stream of cold air. Estimate the maximum capacity of the refrigeration plant required, if it is assumed that the maximum rate of heat extraction from the product is twice the average rate. If the heat-transfer coefficient from the air to the evaporator coils, which form the heat exchanger between the air and the boiling refrigerant, is 22 J m^-2 s^-1°C^-1, calculate the surface area of evaporator coil required if the logarithmic mean temperature drop across the coil is 12°C.

From the tabulated data (Appendix 7) the specific heat of fish is 3.18 kJ kg^-1 °C^-1 above freezing and 1.67 kJ kg^-1°C^-1 below freezing, and the latent heat is 276 kJ/kg^-1.
Enthalpy change in fish: = heat loss above freezing + heat loss below freezing + latent heat
       = (10 x 3.18) + (8 x 1.67) + 276 = 31.8 + 13.4 + 276
                                                   = 321.2 kJ kg^-1
Total heat removed in freezing:
                        15 x 1000 x 321.2 = 4.82 x 10⁶ kJ day^-1
Average rate of heat removal:
            (4.82 x 10⁶)/(24 x 60 x 60) = 55.8 kJ s^-1

If maximum is twice average
then maximum = q = 111.6 kJ s^{-1
= 111.6 x 103}J s^-1

and so                                       A = q/UDT_m
                                            = (111.6 x 10³)/(22 x 12)
                      = 0.42 x 10³ m²
                  = 420 m².

The great advantage of tracing the cycle on the pressure/enthalpy diagram is that from the numerical coordinates of the various cycle points performance parameters can be read off or calculated readily.

EXAMPLE 6.8. Rate of boiling of refrigerant
Ammonia liquid is boiling in an evaporator under an absolute pressure of 120 kPa. Find the temperature and the volumetric rate of evolution of ammonia gas if the heat extraction rate from the surroundings is 300 watts.

From the chart in Appendix 11(b) the boiling temperature of ammonia at a pressure of 120 kPa is -30°C, so this is the evaporator temperature.

Also from the chart, the latent heat of evaporation is:
(enthalpy of saturated vapour - enthalpy of saturated liquid)
= 1.72 - 0.36 = 1.36 MJ kg^{-1 = 1.36 x 103}kJ kg^-1

For a heat-removal rate of 300 watts = 0.3 kJ s^-1 the ammonia evaporation rate is:
(0.3/1360) = 2.2 x 10^-4 kg s^-1.

The chart shows at the saturated vapour point for the cycle (b on Fig. 6.8) that the specific volume of ammonia vapour is:
0.98 m³ kg^-1
and so the volumetric rate of ammonia evolution is:

So far we have been talking of the theoretical cycle. Real cycles differ by, for example, pressure drops in piping, superheating of vapour to the compressor and non-adiabatic compression, but these are relatively minor. Approximations quite good enough for our purposes can be based on the theoretical cycle. If necessary, allowances can be included for particular inefficiencies.

The refrigerant vapour has to be compressed so it can continue round the cycle and be condensed. From the refrigeration demand, the weight of refrigerant required to be circulated can be calculated; each kg s^-1 extracts so many J s^-1 according to the value of (H_b - H_a). From the volume of this refrigerant, the compressor displacement can be calculated, as the compressor has to handle this volume. Because the compressor piston cannot entirely displace all of the working volume of the cylinder, there must be a clearance volume; and because of inefficiencies in valves and ports, the actual amount of refrigerant vapour taken in is less than the theoretical. The ratio of these, actual amount taken in, to the theoretical compressor displacement, is called the volumetric efficiency of the compressor. Both mechanical and volumetric efficiencies can be measured, or taken from manufacturer's data, and they depend on the actual detail of the equipment used.

Consideration of the data from the thermo-dynamic chart and of the refrigeration cycle enables quite extensive calculations to be made about the operation.

EXAMPLE 6.9. Operation of a compressor in a refrigeration system
To meet the requirements of Example 6.7, calculate the speed at which it would be necessary to run a six-cylinder reciprocating ammonia compressor with each cylinder having a 10-cm bore (diameter) and a 12-cm stroke (length of piston travel), assuming a volumetric efficiency of 80%. The condensing temperature is 30°C (determined from the available cooling water temperature) and the evaporating temperature needed is -15°C. Calculate also the theoretical coefficient of performance of this refrigeration system.

From the chart Appendix 11(b), the heat extracted by ammonia at the evaporating temperature of -15°C is (1.74-0.63) = 1.11 MJ kg^{-1 = 1.11x 103}J kg^-1

Under cycle conditions, from chart, the coefficient of performance
= (1.74- 0.63)/(1.97 -1.74) = 1.11/0.23
= 4.8.

Variations in load and in the evaporating or condensing temperatures are often encountered when considering refrigeration systems. Their effects can be predicted by relating them to the basic cycle.

If the heat load increases in the cold store, then the temperature tends to rise and this increases the amount of refrigerant boiling off. If the compressor cannot move this, then the pressure on the suction side of the compressor increases and so the evaporating temperature increases tending to reduce the evaporation rate and correct the situation. However, the effect is to lift the temperature in the cold space and if this is to be prevented additional compressor capacity is required.

As the evaporating pressure, and resultant temperature, change, so the volume of vapour per kilogram of refrigerant changes. If the pressure decreases, this volume increases, and therefore the refrigerating effect, which is substantially determined by the rate of circulation of refrigerant, must also decrease. Therefore if a compressor is required to work from a lower suction pressure its capacity is reduced, and conversely. So at high suction pressures giving high circulation rates, the driving motors may become overloaded because of the substantial increase in quantity of refrigerant circulated in unit time.

Changes in the condenser pressure have relatively little effect on the quantity of refrigerant circulated. However, changes in the condenser pressure and also decreases in suction pressure, have quite a substantial effect on the power consumed per ton of refrigeration. Therefore for an economical plant, it is important to keep the suction pressure as high as possible, compatible with the product requirement for low temperature or rapid freezing, and the condenser pressure as low as possible compatible with the available cooling water or air temperature.

Although in theory a considerable number of fluids might be used in mechanical refrigeration, and historically quite a number including cold air and carbon dioxide have been, those actually in use today are only a very small number. Substantially they include only ammonia, which is used in many large industrial systems, and approved members of a family of halogenated hydrocarbons containing differing proportions of fluorine and chosen according to the particular refrigeration duty required. The reasons for this very small group of practical refrigerants are many.
Important reasons are:

actual vapour pressure/temperature curve for the refrigerant which determines the pressures between which the system must operate for any particular pair of evaporator and condenser temperatures,

refrigerating effect per cubic metre of refrigerant pumped around the system which in turn governs the size of compressors and piping, and the stability and cost of the refrigerant itself.

Ammonia is in most ways the best refrigerant from the mechanical point of view, but its great problem is its toxicity. The thermodynamic chart for ammonia, refrigerant 717 is given in Appendix 11(b).
In working spaces, such as encountered in air conditioning, the halogenated hydrocarbons (often known, after the commercial name, as Freons) are very often used because of safety considerations. They are also used in domestic refrigerators. Environmental problems, in particular due to the effects of any chlorine derivatives and their long lives in the upper atmosphere, have militated heavily, in recent years, against some formerly common halogenated hydrocarbons. Only a very restricted selection has been judged safe. A thermodynamic chart for refrigerant 134a (the reasons for the numbering system are obscure, ammonia being refrigerant 717), the commonest of the safe halogenated hydrocarbon refrigerants, is given in Appendix 11(a). Other charts are available in references such as those provided by the refrigerant manufacturers and in books such as the ASHRAE Guide and Data Books.

Compressors are just basically vapour pumps and much the same types as shown in Fig. 4.3 for liquid pumps are encountered. Their design is highly specialized, particular problems arising from the lower density and viscosity of vapours when compared with liquids. The earliest designs were reciprocating machines with pistons moving horizontally or vertically, at first in large cylinders and at modest speeds, and then increasingly at higher speeds in smaller cylinders. An important aspect in compressor choice is the compression ratio, being the ratio of the absolute pressure of discharge from the compressor, to the inlet suction pressure. Reciprocating compressors can work effectively at quite high compression ratios (up to 6 or 7 to 1). Higher overall compression ratios are best handled by putting two or more compressors in series and so sharing the overall compression ratio between them.

For smaller compression ratios and for handling the large volumes of vapours encountered at low temperatures and pressures, rotary vane compressors are often used, and for even larger volumes, centrifugal compressors, often with many stages, can be used. A recent popular development is the screw compressor, analogous to a gear pump, which has considerable flexibility. Small systems are often "hermetic", implying that the motor and compressor are sealed into one casing with the refrigerant circulating through both. This avoids rotating seals through which refrigerant can leak. The familiar and very dependable units in household refrigerators are almost universally of this type.

The evaporator is the only part of the refrigeration equipment that enters directly into food processing operations. Heat passes from the food to the heat-transfer medium, which may be air or liquid, thence to the evaporator and so to the refrigerant. Thermal coupling is in some cases direct, as in a plate freezer. In this, the food to be frozen is placed directly on or between plates, within which the refrigerant circulates. Another familiar example of direct thermal coupling is the chilled slab in a shop display.

Generally, however, the heat transfer medium is air, which moves either by forced or natural circulation between the heat source, the food and the walls warmed by outside air, and the heat sink which is the evaporator. Sometimes the medium is liquid such as in the case of immersion freezing in propylene glycol or in alcohol-water mixtures. Then there are some cases in which the refrigerant is, in effect, the medium such as immersion or spray with liquid nitrogen. Sometimes there is also a further intermediate heat-transfer medium, so as to provide better control, or convenience, or safety. An example is in some milk chillers, where the basic refrigerant is ammonia; this cools glycol, which is pumped through a heat exchanger where it cools the milk. A sketch of some types of evaporator system is given in Fig. 6.10.

In the freezer or chiller, the heat transfer rates both from the food and to the evaporator, depend upon fluid or gas velocities and upon temperature differences. The values for the respective heat transfer coefficients can be estimated by use of the standard heat transfer relationships. Typical values that occur in freezing equipment are given in Table 6.3.

	J m^-2 s^-1 °C^-1
Still-air freezing (including radiation to coils)	9
Still-air freezing (no radiation)	6
Air-blast freezing 3 m s^-1	18
Air-blast freezing 5 m s^-1	30
Liquid-immersion freezing	600
Plate freezing	120

Temperature differences across evaporators are generally of the order 3 - 10°C. The calculations for heat transfer can be carried out using the methods which have been discussed in other sections, including their relationship to freezing and freezing times, as the evaporators are just refrigerant-to-air heat exchangers.

The evaporator surfaces are often extended by the use of metal fins that are bonded to the evaporator pipe surface. The reason for this construction is that the relatively high metal conductance, compared with the much lower surface conductance from the metal surface to the air, maintains the fin surface substantially at the coil temperature. A slight rise in temperature along the fin can be accounted for by including in calculations a fin efficiency factor. The effective evaporator area is then calculated by the relationship

where A is the equivalent total evaporator surface area, A_p is the coil surface area, called the primary surface, A_s is the fin surface area, called the secondary surface and f (phi) is the fin efficiency.
Values of f lie between 65% and 95% in the usual designs, as shown for example in DKV Arbeitsblatt 2-02, 1950.

Chilling of foods is a process by which their temperature is reduced to the desired holding temperature just above the freezing point of food, usually in the region of -2 to 2°C

Many commercial chillers operate at higher temperatures, up to 10-12°C. The effect of chilling is only to slow down deterioration changes and the reactions are temperature dependent. So the time and temperature of holding the chilled food determine the storage life of the food.

Rates of chilling are governed by the laws of heat transfer which have been described in previous sections. It is an example of unsteady-state heat transfer by convection to the surface of the food and by conduction within the food itself. The medium of heat exchange is generally air, which extracts heat from the food and then gives it up to refrigerant in the evaporator. As explained in the heat-transfer section, rates of convection heat transfer from the surface of food and to the evaporator are much greater if the air is in movement, being roughly proportional to v^0.8.

To calculate chilling rates it is therefore necessary to evaluate:
(a) surface heat transfer coefficient,
(b) resistance offered to heat flow by any packaging material that may be placed round the food,
(c) appropriate unsteady state heat conduction equation.

Although the shapes of most foodstuffs are not regular, they often approximate the shapes of slabs, bricks, spheres and cylinders.

EXAMPLE 6.10. Chilling of fresh apples
Before apples are loaded into a cool store, it is wished to chill them to a central temperature of 5°C so as to avoid problems of putting warm apples with the colder ones in storage. The apples, initially at 25°C, are considered to be spheres of 7 cm diameter and the chilling is to be carried out using air at -1°C and at a velocity which provides a surface heat-transfer coefficient of 30 J m^-2 s^-1 °C^-1. The physical properties of the apples are k = 0.5 J m^-1 s^-1 °C^-1, r = 930 kg m^-3, c = 3.6 kJ kg^-1 °C^-1. Calculate the time necessary to chill the apples so that their centres reach 5°C.

This is an example of unsteady-state cooling and can be solved by application of Fig. 5.3,

                      so t = [0.46x 3600 x 930 x (0.035)²]/0.5
                             = 3773 s
                             = 1.05 h

A full analysis of chilling must, in addition to heat transfer, take mass transfer into account if the food surfaces are moist and the air is unsaturated. This is a common situation and complicates chilling analysis.

Water makes up a substantial proportion of almost all foodstuffs and so freezing has a marked physical effect on the food. Because of the presence of substances dissolved in the water, food does not freeze at one temperature but rather over a range of temperatures. At temperatures just below the freezing point of water, crystals that are almost pure ice form in the food and so the remaining solutions become more concentrated. Even at low temperatures some water remains unfrozen, in very concentrated solutions.

In the freezing process, added to chilling is the removal of the latent heat of freezing. This latent heat has to be removed from any water that is present. Since the latent heat of freezing of water is 335 kJ kg^-1, this represents the most substantial thermal quantity entering into the process. There may be other latent heats, for example the heats of solidification of fats which may be present, and heats of solution of salts, but these are of smaller magnitude than the latent heat of freezing of water. Also the fats themselves are seldom present in foods in as great a proportion as water.

Because of the latent-heat-removal requirement, the normal unsteady-state equations cannot be applied to the freezing of foodstuffs. The coefficients of heat transfer can be estimated by the following equation:

where h_s is the total surface heat-transfer coefficient, h_c is the convection heat transfer coefficient, x is the thickness of packing material, k is the thermal conductivity of the packing material and h_r is the radiation heat-transfer coefficient.

A full analytical solution of the rate of freezing of food cannot be obtained. However, an approximate solution, due to Plank, is sufficient for many practical purposes. Plank assumed that the freezing process:
(a) commences with all of the food unfrozen but at its freezing temperature,
(b) occurs sufficiently slowly for heat transfer in the frozen layer to take place under steady-state conditions.

Making these assumptions, freezing rates for bodies of simple shapes can be calculated. As an example of the method, the time taken to freeze to the centre of a slab whose length and breadth are large compared with the thickness, will be calculated.

Rates of heat transfer are equal from either side of the slab. Assume that at time t a thickness x of the slab of area A has been frozen as shown in Fig. 6.11. The temperature of the freezing medium is T_a. The freezing temperature of the foodstuff is T, and the surface temperature of the food is T_s. The thermal conductivity of the frozen food is k, l is the latent heat of the foodstuff and r is its density.

The rate of movement of the freezing boundary multiplied by the latent heat equals the rate of heat transfer of heat to the boundary:

Now, all of the heat removed at the freezing boundary must be transmitted to the surface through the frozen layer; if the frozen layer is in the steady-state condition we have:

Similarly, this quantity of heat must be transferred to the cooling medium from the food surface, so:
q = Ah_s(T_s - T_a)

Since the same heat flow passing through the surface also passes through the frozen layer and is removed from the water as it freezes in the centre of the block:

Now, if the thickness of the slab is a, the time taken for the centre of the slab at x = a/2 to freeze can be obtained by integrating from x = 0, to x = a/2 during which time t goes from t to t_f .

In his papers Plank (1913, 1941) derived his equation in more general terms and found that for brick-shaped solids the change is in numerical terms, called shape factors, only.

where P = 1/2, R = 1/8 for a slab; P = 1/4, R = 1/16 for an infinitely long cylinder, and P = 1/6, R = 1/24 for a cube or a sphere. Brick-shaped solids have values of P and R lying between those for slabs and those for cubes. Appropriate values of P and R for a brick-shaped solid can be obtained from the graph in Fig. 6.12. In this figure, b₁ and b₂ are the ratios of the two longest sides to the shortest. It does not matter in what order they are taken.

Figure 6.12 Coefficients in Plank’s Equation
Adapted from Ede, Modern Refrigeration, 1952

Because the assumptions made in the derivation of Plank's equation lead to errors, which tend towards under-estimation of freezing times, more accurate predictions can be made if some allowances are made for this. One step is to use the total enthalpy changes from the initial to the final state of the product being frozen, that is to include the sensible heat changes both above and below the freezing temperature in addition to the latent heat. Even with this addition the prediction will still be about 20% or so lower than equivalent experimental measurements for brick shapes indicate. Adaptations of Plank's equation have been proposed which correspond better with experimental results, such as those in Cleland and Earle (1982), but they are more complicated.

EXAMPLE 6.11. Freezing of a slab of meat
If a slab of meat is to be frozen between refrigerated plates with the plate temperature at -34°C, how long will it take to freeze if the slab is 10 cm thick and the meat is wrapped in cardboard 1 mm thick on either side of the slab? What would be the freezing time if the cardboard were not present? Assume that for the plate freezer, the surface heat-transfer coefficient is 600 J m^-2 s^-1 °C^-1, the thermal conductivity of cardboard is 0.06 J m^-1 s^-1° C^-1 the thermal conductivity of frozen meat is 1.6 J m^-1 s^-1 °C^-1, its latent heat is 2.56 x 10⁵ J kg^-1 and density 1090 kg m^-3. Assume also that meat freezes at -2°C.

In a plate freezer, the thickness of the slab is the only dimension that is significant. The case can be treated as equivalent to an infinite slab, and therefore the constants in Plank's equation are 1/2 and 1/8.

This estimate can be improved by adding to the latent heat of freezing, the enthalpy change above the freezing temperature and below the freezing temperature. Above freezing, assume the meat with a specific heat of 3.22 x 10³ J kg^-1°C^-1 starts at +10°C and goes to -2°C, needing 3.9 x 10⁴ J kg^-1. Below the freezing temperature assume the meat goes from -2°C to the mean of -2°C and -34°C, that is -18°C, with a specific heat of 1.67 x 10³ J kg^-1°C^-1, needing 2.8 x 10⁴ J kg^-1. This, with the latent heat of 2.56 x 10⁵ Jkg, gives a total of 3.23 x 10⁵ J kg^-1 and amended freezing times of

If instead of a slab, cylinder, or cube, the food were closer to a brick shape then an additional 20% should be added.

Thus, with a knowledge of the thermal constants of a foodstuff, required freezing times can be estimated by the use of Plank's equation. Appendix 7 gives values for the thermal conductivities, and the latent heats and densities, of some common foods.

The analysis using Plank's equation separates the total freezing time effectively into two terms, one intrinsic to the food material to be frozen, and the other containing the surface heat transfer coefficient which can be influenced by the process equipment. Therefore for a sensitivity analysis, it can be helpful to write Plank's equation in dimensionless form, substituting DH for lr:

in which h (eta) can be regarded as an efficiency of coupling of the freezing medium to the food varying from 1 for Bi ® ¥, to 0 for Bi ® 0. Taking an intrinsic freezing time t_f’ for the case of unit driving force, DT = 1; then for general DT :
t_f = t_f' /(hDT) (6.9)

and (6.9) can be used very readily to examine the influence of freezing medium temperature, and of surface heat transfer coefficient through the Biot number, on the actual freezing time. Thus the sensitivity of freezing time to process variations can be taken quickly into account.

EXAMPLE 6.12. Freezing time of a carton of meat: controllable factors
Determine the intrinsic freezing time for the carton of Example 6.11. By putting the equation for the freezing time in the form of eqn. (6.9) evaluate the effect of
(a) changes in the temperature of the plates to -20°C, - 25°C, and -30°C
(b) effect of doubling the thickness of the cardboard and
(c) effect of decreased surface coefficients due to poor contact which drops the surface heat transfer coefficient to 100 J m^-2 s^-1°C^-1

P/R = 4, and so h = 3.3/7.3 = 0.45
And from Example 6.11, calculated freezing time is 4.2 h with driving force 32°C
Therefore t_f‘ = 4.2 x 0.45 x 32 = 60.5 h,

for (a) t_f = 60.5 /(0.45 x 18) = 7.5 h for -20°C, 60.5/(0.45 x 23) = 5.8 h for -25°C, and 60.5/(0.45 x 28) = 4.8 h for -30°C

for (b) x/k becomes 0.034, therefore 1/h_s = 0.034+0.0017
                h_{s =}28 J m^-2 s^-1°C^-1
Bi becomes (28 x 0.1)/1.6 = 1.75
Therefore   h becomes 0.30
so              t_f = 60.5 x 1/0.30 + 1/32
                    = 6.3 h

for (c) 1/ h_s becomes 0.017 + 0.01 = 0.027,
              h_s = 37 J m^-2 s^-1°C^-1
Bi becomes (37 x 0.1)/1.6 = 2.3
Therefore   h becomes 0.37
so              t_f = 60.5 x 1/32 x 1/0.37
                    = 5.1 h

For cold storage, the requirement for refrigeration comes from the need to remove the heat:

coming into the store from the external surroundings through insulation

from sources within the store such as motors, lights and people (each worker contributing something of the order of 0.5 kW)

from the foodstuffs.

Heat penetrating the walls can be estimated, knowing the overall heat-transfer coefficients including the surface terms and the conductances of the insulation, which may include several different materials. The other heat sources require to be considered and summed. Detailed calculations can be quite complicated but for many purposes simple methods give a reasonable estimate.

Heat-Transfer Applications > SUMMARY, PROBLEMS