Heat capacity of flue gases at different temperatures table. Physical properties of air, gases and water. "Samara State Technical University"
Humid air is a mixture of dry air and water vapor. In unsaturated air, moisture is in the state of superheated vapor, and therefore the properties of moist air can be approximately described by the laws of ideal gases.
The main characteristics of humid air are:
1. Absolute humidity g, which determines the amount of water vapor contained in 1 m 3 of moist air. Water vapor occupies the entire volume of the mixture, so the absolute humidity of the air is equal to the mass of 1 m 3 of water vapor or vapor density, kg / m 3
2. Relative humidity j is expressed by the ratio of the absolute humidity of the air to its maximum possible humidity at the same pressure and temperature, or by the ratio of the mass of water vapor contained in 1 m 3 of moist air to the mass of water vapor required to completely saturate 1 m 3 of moist air at the same pressure and temperature.
Relative humidity determines the degree of saturation of air with moisture:
, (1.2)
where is the partial pressure of water vapor corresponding to its density Pa; - pressure of saturated steam at the same temperature, Pa; - the maximum possible amount of steam in 1 m 3 of saturated moist air, kg / m 3; - vapor density at its partial pressure and temperature of moist air, kg/m 3 .
Relation (1.2) is valid only when it can be assumed that the liquid vapor is an ideal gas up to the saturation state.
The density of moist air r is the sum of the densities of water vapor and dry air at partial pressures of 1 m 3 of moist air at a temperature of humid air T, TO:
(1.3)
where is the density of dry air at its partial pressure of 1 m 3 of moist air, kg / m 3; - partial pressure of dry air, Pa; - gas constant of dry air, J/(kg×K).
Expressing and by the equation of state for air and water vapor, we obtain
, (1.5)
where is the mass flow rate of air and water vapor, kg/s.
These equalities are valid for the same volume V humid air at the same temperature. Dividing the second equality by the first, we get another expression for the moisture content
. (1.6)
Substituting here the values of gas constants for air J/(kg×K) and for water vapor J/(kg×K), we obtain the value of moisture content expressed in kilograms of water vapor per 1 kg of dry air
. (1.7)
Replacing the partial air pressure with the value , where from the previous and AT is the barometric air pressure in the same units as R, we get for moist air under barometric pressure
. (1.8)
Thus, at a given barometric pressure, the moisture content of air depends only on the partial pressure of water vapor. The maximum possible moisture content in the air, from where
. (1.9)
Since the saturation pressure increases with temperature, the maximum possible amount of moisture that can be contained in the air depends on its temperature, and the more, the higher the temperature. If equations (1.7) and (1.8) are solved for and , then we obtain
(1.10)
. (1.11)
The amount of moist air in cubic meters per 1 kg of dry air is calculated by the formula
(1.12)
Specific volume of moist air v, m 3 / kg, is determined by dividing the volume of moist air by the mass of the mixture per 1 kg of dry air:
Humid air as a heat carrier is characterized by an enthalpy (in kilojoules per 1 kg of dry air) equal to the sum of the enthalpies of dry air and water vapor
(1.14)
where is the specific heat capacity of dry air, kJ/(kg×K); t– air temperature, °C; i- enthalpy of superheated steam, kJ/kg.
The enthalpy of 1 kg of dry saturated water vapor at low pressures is determined by the empirical formula, kJ/kg:
where is a constant coefficient approximately equal to the enthalpy of steam at a temperature of 0 °C; = 1.97 kJ/(kg×K) – specific heat capacity of steam.
Substituting the values i into expression (1.14) and taking the specific heat of dry air constant and equal to 1.0036 kJ / (kg × K), we find the enthalpy of moist air in kilojoules per 1 kg of dry air:
Equations similar to those discussed above are used to determine the wet gas parameters.
, (1.17)
where is the gas constant for the test gas; R- gas pressure.
Gas enthalpy, kJ/kg,
where is the specific heat capacity of the gas, kJ/(kg×K).
Absolute moisture content of gas:
. (1.19)
When calculating contact heat exchangers for air-water heat carriers, you can use the data in Table. 1.1-1.2 or calculated dependencies for determining the physicochemical parameters of air (1.24-1.34) and water (1.35). For flue gases, the data in Table 1 can be used. 1.3.
Wet gas density, kg / m 3:
, (1.20)
where is the density of dry gas at 0 ° C, kg / m 3; M g, M p are the molecular masses of gas and vapor.
Wet gas dynamic viscosity coefficient, Pa×s:
, (1.21)
where is the coefficient of dynamic viscosity of water vapor, Pa×s; - coefficient of dynamic viscosity of dry gas, Pa×s; 
- mass concentration of steam, kg/kg.
Specific heat capacity of wet gas, kJ/(kg×K):
Wet gas thermal conductivity coefficient, W/(m×K):
, (1.23)
where k is the adiabatic index; AT– coefficient (for monatomic gases AT= 2.5; for diatomic gases AT= 1.9; for triatomic gases AT = 1,72).
Table 1.1. Physical Properties dry air ( R= 0.101 MPa)
| t, °C | , kg / m 3 | , kJ/(kg×K) | , W/(m×K) | , Pa×s | , m 2 /s | Pr |
| -20 | 1,395 | 1,009 | 2,28 | 16,2 | 12,79 | 0,716 |
| -10 | 1,342 | 1,009 | 2,36 | 16,7 | 12,43 | 0,712 |
| 1,293 | 1,005 | 2,44 | 17,2 | 13,28 | 0,707 | |
| 1,247 | 1,005 | 2,51 | 17,6 | 14,16 | 0,705 | |
| 1,205 | 1,005 | 2,59 | 18,1 | 15,06 | 0,703 | |
| 1,165 | 1,005 | 2,67 | 18,6 | 16,00 | 0,701 | |
| 1,128 | 1,005 | 2,76 | 19,1 | 16,96 | 0,699 | |
| 1,093 | 1,005 | 2,83 | 19,6 | 17,95 | 0,698 | |
| 1,060 | 1,005 | 2,90 | 20,1 | 18,97 | 0,696 | |
| 1,029 | 1,009 | 2,96 | 20,6 | 20,02 | 0,694 | |
| 1,000 | 1,009 | 3,05 | 21,1 | 21,09 | 0,692 | |
| 0,972 | 1,009 | 3,13 | 21,5 | 22,10 | 0,690 | |
| 0,946 | 1,009 | 3,21 | 21,9 | 23,13 | 0,688 | |
| 0,898 | 1,009 | 3,34 | 22,8 | 25,45 | 0,686 | |
| 0,854 | 1,013 | 3,49 | 23,7 | 27,80 | 0,684 | |
| 0,815 | 1,017 | 3,64 | 24,5 | 30,09 | 0,682 | |
| 0,779 | 1,022 | 3,78 | 25,3 | 32,49 | 0,681 | |
| 0,746 | 1,026 | 3,93 | 26,0 | 34,85 | 0,680 | |
| 0,674 | 1,038 | 4,27 | 27,4 | 40,61 | 0,677 | |
| 0,615 | 1,047 | 4,60 | 29,7 | 48,33 | 0,674 | |
| 0,566 | 1,059 | 4,91 | 31,4 | 55,46 | 0,676 | |
| 0,524 | 1,068 | 5,21 | 33,6 | 63,09 | 0,678 | |
| 0,456 | 1,093 | 5,74 | 36,2 | 79,38 | 0,687 | |
| 0,404 | 1,114 | 6,22 | 39,1 | 96,89 | 0,699 | |
| 0,362 | 1,135 | 6,71 | 41,8 | 115,4 | 0,706 | |
| 0,329 | 1,156 | 7,18 | 44,3 | 134,8 | 0,713 | |
| 0,301 | 1,172 | 7,63 | 46,7 | 155,1 | 0,717 | |
| 0,277 | 1,185 | 8,07 | 49,0 | 177,1 | 0,719 | |
| 0,257 | 1,197 | 8,50 | 51,2 | 199,3 | 0,722 | |
| 0,239 | 1,210 | 9,15 | 53,5 | 233,7 | 0,724 |
Thermophysical properties dry air can be approximated by the following equations.
Kinematic viscosity of dry air at temperatures from -20 to +140 ° C, m 2 / s:
Pa; (1.24)
and from 140 to 400 °С, m2/s:
. (1.25)
Table 1.2. Physical properties of water in a state of saturation
| t, °C | , kg / m 3 | , kJ/(kg×K) | , W/(m×K) | , m 2 /s | , N/m | Pr | |
| 999,9 | 4,212 | 55,1 | 1,789 | -0,63 | 756,4 | 13,67 | |
| 999,7 | 4,191 | 57,4 | 1,306 | 0,7 | 741,6 | 9,52 | |
| 998,2 | 4,183 | 59,9 | 1,006 | 1,82 | 726,9 | 7,02 | |
| 995,7 | 4,174 | 61,8 | 0,805 | 3,21 | 712,2 | 5,42 | |
| 992,2 | 4,174 | 63,5 | 0,659 | 3,87 | 696,5 | 4,31 | |
| 988,1 | 4,174 | 64,8 | 0,556 | 4,49 | 676,9 | 3,54 | |
| 983,2 | 4,179 | 65,9 | 0,478 | 5,11 | 662,2 | 2,98 | |
| 977,8 | 4,187 | 66,8 | 0,415 | 5,70 | 643,5 | 2,55 | |
| 971,8 | 4,195 | 67,4 | 0,365 | 6,32 | 625,9 | 2,21 | |
| 965,3 | 4,208 | 68,0 | 0,326 | 6,95 | 607,2 | 1,95 | |
| 958,4 | 4,220 | 68,3 | 0,295 | 7,52 | 588,6 | 1,75 | |
| 951,0 | 4,233 | 68,5 | 0,272 | 8,08 | 569,0 | 1,60 | |
| 943,1 | 4,250 | 68,6 | 0,252 | 8,64 | 548,4 | 1,47 | |
| 934,8 | 4,266 | 68,6 | 0,233 | 9,19 | 528,8 | 1,36 | |
| 926,1 | 4,287 | 68,5 | 0,217 | 9,72 | 507,2 | 1,26 | |
| 917,0 | 4,313 | 68,4 | 0,203 | 10,3 | 486,6 | 1,17 | |
| 907,4 | 4,346 | 68,3 | 0,191 | 10,7 | 466,0 | 1,10 | |
| 897,3 | 4,380 | 67,9 | 0,181 | 11,3 | 443,4 | 1,05 | |
| 886,9 | 4,417 | 67,4 | 0,173 | 11,9 | 422,8 | 1,00 | |
| 876,0 | 4,459 | 67,0 | 0,165 | 12,6 | 400,2 | 0,96 | |
| 863,0 | 4,505 | 66,3 | 0,158 | 13,3 | 376,7 | 0,93 |
Wet gas density, kg/m3.
The thermophysical properties of gaseous combustion products necessary for calculating the dependence of various parameters on the temperature of a given gaseous medium can be established on the basis of the values given in the table. In particular, these dependences for the heat capacity are obtained in the form:
C psm = a -1/ d,
where a = 1,3615803; b = 7,0065648; c = 0,0053034712; d = 20,761095;
C psm = a + bT sm + cT 2 sm,
where a = 0,94426057; b = 0,00035133267; c = -0,0000000539.
The first dependence is preferable in terms of approximation accuracy, the second dependence can be taken to carry out calculations of lower accuracy.
Physical parameters flue gases
(at P = 0.0981 MPa; R CO2 = 0.13; p H2O = 0.11; R N2 = 0.76)
| t, °С | γ, N m -3 | with p, W (m 2 ° С) -1 | λ 10 2, W (m K) -1 | a 10 6, m 2 s -1 | μ 10 6 , Pa s | v 10 6, m 2 s -1 | Pr |
| 12,704 | 1,04 | 2,28 | 16,89 | 15,78 | 12,20 | 0,72 | |
| 9,320 | 1,07 | 3,13 | 30,83 | 20,39 | 21,54 | 0,69 | |
| 7,338 | 1,10 | 4,01 | 48,89 | 24,50 | 32,80 | 0,67 | |
| 6,053 | 1,12 | 4,84 | 69,89 | 28,23 | 45,81 | 0,65 | |
| 5,150 | 1,15 | 5,70 | 94,28 | 31,69 | 60,38 | 0,64 | |
| 4,483 | 1,18 | 6,56 | 121,14 | 34,85 | 76,30 | 0,63 | |
| 3,973 | 1,21 | 7,42 | 150,89 | 37,87 | 93,61 | 0,62 | |
| 3,561 | 1,24 | 8,27 | 183,81 | 40,69 | 112,10 | 0,61 | |
| 3,237 | 1,26 | 9,15 | 219,69 | 43,38 | 131,80 | 0,60 | |
| 2,953 | 1,29 | 10,01 | 257,97 | 45,91 | 152,50 | 0,59 | |
| 2,698 | 1,31 | 10,90 | 303,36 | 48,36 | 174,30 | 0,58 | |
| 2,521 | 1,32 | 11,75 | 345,47 | 40,90 | 197,10 | 0,57 | |
| 2,354 | 1,34 | 12,62 | 392,42 | 52,99 | 221,00 | 0,56 |
APPENDIX 3
(reference)
Air and smoke permeability of air ducts and valves
1. To determine leaks or air leaks in relation to ventilation ducts smoke control systems, the following formulas can be used, obtained by approximation of tabular data:
for class H air ducts (in the pressure range 0.2 - 1.4 kPa): ΔL = a(R - b)With, where ΔL- suctions (leaks) of air, m 3 / m 2 h; R- pressure, kPa; a = 10,752331; b = 0,0069397038; With = 0,66419906;
for class P air ducts (in the pressure range 0.2 - 5.0 kPa): where a = 0,00913545; b=-3.1647682 10 8 ; c =-1.2724412 10 9 ; d= 0,68424233.
2. For normally closed fire dampers, the numerical values of the specific characteristic of resistance to smoke and gas permeation depending on the gas temperature correspond to the data obtained during bench fire tests of various products at the experimental base of VNIIPO:
| 1. General provisions. 2 2. Initial data. 3 3. Exhaust smoke ventilation. 4 3.1. Removal of combustion products directly from the burning room. 4 3.2. Removal of combustion products from adjacent premises. 7 4. Supply smoke ventilation. 9 4.1. Air supply to stairwells. 9 4.2. Air supply to lift shafts.. 14 4.3. Air supply to the vestibule locks.. 16 4.4. Compensating air supply. 17 5. Specifications equipment. 17 5.1. Equipment for exhaust smoke ventilation systems. 17 5.2. Equipment for supply smoke ventilation systems. 21 6. Fire control modes. 21 References.. 22 Appendix 1. Determination of the main parameters of the fire load of premises. 22 Appendix 2. Thermophysical properties of flue gases. 24 Appendix 3. Air and smoke permeability of air ducts and valves. 25 |
State educational institution higher professional education
"Samara State Technical University"
Department of Chemical Technology and Industrial Ecology
COURSE WORK
in the discipline "Technical thermodynamics and heat engineering"
Topic: Calculation of the installation for the recovery of heat from the waste gases of a process furnace
Completed by: Student Ryabinina E.A.
ZF course III group 19
Checked by: Consultant Churkina A.Yu.
Samara 2010
Introduction
Most chemical enterprises generate high- and low-temperature thermal waste, which can be used as secondary energy resources (SER). These include flue gases from various boilers and process furnaces, cooled streams, cooling water and exhaust steam.
Thermal VER to a large extent cover the heat demand of individual industries. Thus, in the nitrogen industry, more than 26% of the heat demand is met by means of VER, in the soda industry - more than 11%.
The number of used HORs depends on three factors: the temperature of the HORs, their thermal power and the continuity of the output.
At present, the most widespread is the utilization of heat from industrial waste gases, which have a high temperature potential for almost all fire engineering processes and can be used continuously in most industries. Waste gas heat is the main component of the energy balance. It is used mainly for technological, and in some cases - for energy purposes (in waste heat boilers).
However, the widespread use of high-temperature thermal VERs is associated with the development of methods for utilization, including the heat of hot slags, products, etc., new methods for utilizing the heat of exhaust gases, as well as with the improvement of the designs of existing utilization equipment.
1. Description technological scheme
In tube furnaces without a convection chamber, or in radiant convection type furnaces, but having a relatively high initial temperature of the heated product, the temperature of the exhaust gases can be relatively high, which leads to increased heat loss, reduced furnace efficiency and higher fuel consumption. Therefore, it is necessary to use the heat of the waste gases. This can be achieved either by using an air heater that heats the air entering the furnace for fuel combustion, or by installing waste heat boilers that make it possible to obtain water vapor necessary for technological needs.
However, for the implementation of air heating, additional costs are required for the construction of an air heater, blowers, as well as additional power consumption consumed by the blower motor.
To ensure the normal operation of the air heater, it is important to prevent the possibility of corrosion of its surface from the side of the flue gas flow. This phenomenon is possible when the temperature of the heat exchange surface is lower than the dew point temperature; at the same time, part of the flue gases, directly in contact with the surface of the air heater, is significantly cooled, the water vapor contained in them partially condenses and, absorbing sulfur dioxide from the gases, forms an aggressive weak acid.
The dew point corresponds to the temperature at which the saturated vapor pressure of water is equal to the partial pressure of water vapor contained in the flue gases.
One of the most reliable ways to protect against corrosion is to preheat the air in some way (for example, in water or steam heaters) to a temperature above the dew point. Such corrosion can also occur on the surface of convection pipes if the temperature of the raw material entering the furnace is below the dew point.
The source of heat to increase the temperature of saturated steam is the oxidation reaction (combustion) of the primary fuel. The flue gases formed during combustion give off their heat in the radiation and then convection chambers to the raw material flow (steam). Superheated water vapor enters the consumer, and the combustion products leave the furnace and enter the waste heat boiler. At the outlet of the KU, saturated water vapor is fed back to the steam superheating furnace, and the flue gases, being cooled by the feed water, enter the air heater. From the air heater, the flue gases enter the CTAN, where the water flowing through the coil is heated and goes directly to the consumer, and the flue gases are released into the atmosphere.
2. Furnace calculation
2.1 Calculation of the combustion process
Let's determine the lower calorific value of fuel combustion Q R n. If the fuel is an individual hydrocarbon, then its calorific value Q R n equal to the standard heat of combustion minus the heat of vaporization of water in the combustion products. It can also be calculated from the standard thermal effects of the formation of initial and final products based on the Hess law.
For a fuel consisting of a mixture of hydrocarbons, the calorific value is determined according to the additivity rule:
where Q pi n- heat of combustion i-th fuel component;
y i- concentration i-th fuel component in fractions of a unit, then:
Q R n cm = 35.84 ∙ 0.987 + 63.80 ∙ 0.0033+ 91.32 ∙ 0.0012+ 118.73 ∙ 0.0004 + 146.10 ∙ 0.0001 \u003d 35.75 MJ / m 3.
Molar mass of fuel:
M m = Σ M i ∙ y i ,
where M i- molar mass i-th fuel component, from here:
M m = 16.042 ∙ 0.987 + 30.07 ∙ 0.0033 + 44.094 ∙ 0.0012 + 58.120 ∙ 0.0004 + 72.15 ∙ 0.0001 + 44.010 ∙ 0.001+ 28.01 ∙ 0.002 = 28.01 ∙ 0.002
kg / m 3,
then Q R n cm, expressed in MJ/kg, is equal to:
MJ/kg.
The calculation results are summarized in Table. one:
Fuel composition Table 1
Let us determine the elemental composition of the fuel, % (mass):
,
where n i C , n i H , n i N , n i O- the number of carbon, hydrogen, nitrogen and oxygen atoms in the molecules of individual components that make up the fuel;
The content of each component of the fuel, wt. %;
x i- the content of each component of the fuel, they say. %;
M i is the molar mass of the individual fuel components;
M m is the molar mass of the fuel.
Composition check :
C + H + O + N = 74.0 + 24.6 + 0.2 + 1.2 = 100% (mass).
Let us determine the theoretical amount of air required to burn 1 kg of fuel; it is determined from the stoichiometric equation of the combustion reaction and the oxygen content in the atmospheric air. If the elemental composition of the fuel is known, the theoretical amount of air L0, kg/kg, is calculated by the formula:
In practice, to ensure the completeness of fuel combustion, an excess amount of air is introduced into the furnace, we find the actual air flow at α = 1.25:
L = aL 0 ,
where L- actual air consumption;
α - coefficient of excess air,
L = 1.25∙17.0 = 21.25 kg/kg.
Specific air volume (n.a.) for combustion of 1 kg of fuel:
where ρ in= 1.293 - air density under normal conditions,
m 3 / kg.
Let's find the amount of combustion products formed during the combustion of 1 kg of fuel:
if the elemental composition of the fuel is known, then the mass composition of flue gases per 1 kg of fuel during its complete combustion can be determined on the basis of the following equations:
where mCO2 , mH2O , mN2 , mO2- mass of the corresponding gases, kg.
Total amount of combustion products:
m p. s = m CO2 + m H2O + m N2 + m O2 ,
m p. s= 2.71 + 2.21 + 16.33 + 1.00 = 22.25 kg/kg.
Checking the received value:
where W f- specific consumption of injector steam during liquid fuel combustion, kg/kg (for gas fuel W f = 0),
Since the fuel is a gas, we neglect the moisture content in the air and do not take into account the amount of water vapor.
Let us find the volume of combustion products under normal conditions formed during the combustion of 1 kg of fuel:
where m i- the mass of the corresponding gas formed during the combustion of 1 kg of fuel;
ρi- the density of this gas under normal conditions, kg / m 3;
M i is the molar mass of the given gas, kg/kmol;
22.4 - molar volume, m 3 / kmol,
m 3 /kg; 
m 3 /kg;
m 3 /kg; 
m 3 / kg.
The total volume of combustion products (n.a.) at the actual air flow:
V = V CO2 + V H2O + V N2 + V O2 ,
V = 1.38 + 2.75 + 13.06 + 0.70 \u003d 17.89 m 3 / kg.
Density of combustion products (n.a.):
kg / m 3.
Let us find the heat capacity and enthalpy of combustion products of 1 kg of fuel in the temperature range from 100 °C (373 K) to 1500 °C (1773 K) using the data in Table. 2.
Average specific heat capacities of gases with p, kJ/(kg∙K) table 2
| t, °С |
|||||
The enthalpy of flue gases generated during the combustion of 1 kg of fuel:
where with CO2 , with H2O , with N2 , with O2- average specific heat capacities at constant pressure of the corresponding lawn at temperature t, kJ/(kg K);
with t is the average heat capacity of flue gases generated during the combustion of 1 kg of fuel at a temperature t, kJ/(kg K);
at 100 °С: kJ/(kg∙K);
at 200 °С: kJ/(kg∙K);
at 300 °C: kJ/(kg∙K);
at 400 °С: kJ/(kg∙K);
at 500 °С: kJ/(kg∙K);
at 600 °C: kJ/(kg∙K);
at 700 °С: kJ/(kg∙K);
at 800 °С: kJ/(kg∙K);
at 1000 °С: kJ/(kg∙K);
at 1500 °C: kJ/(kg∙K);
The results of the calculations are summarized in Table. 3.
Enthalpy of combustion products Table 3
According to Table. 3 build a dependency graph H t = f ( t ) (Fig. 1) see Attachment .
2.2 Calculation heat balance furnace, furnace efficiency and fuel consumption
Heat flux taken up by water vapor in the furnace (useful heat load):
where G- the amount of superheated water vapor per unit time, kg/s;
H vp1 and H vp2
We take the temperature of the outgoing flue gases to be 320 °C (593 K). Heat loss by radiation in environment will be 10%, with 9% of them lost in the radiant chamber, and 1% in the convection chamber. Furnace efficiency η t = 0.95.
Heat losses from chemical underburning, as well as the amount of heat of the incoming fuel and air, are neglected.
Let's determine the efficiency of the furnace:
where uh is the enthalpy of the combustion products at the temperature of the flue gases leaving the furnace, t uh; the temperature of the outgoing flue gases is usually assumed to be 100 - 150 ° C higher than the initial temperature of the raw material at the furnace inlet; q sweat- heat loss by radiation to the environment, % or fraction of Q floor ;
Fuel consumption, kg/s:
kg/s.
2.3 Calculation of the radiant chamber and convection chamber
We set the flue gas temperature at the pass: t P\u003d 750 - 850 ° С, we accept
t P= 800 °C (1073 K). Enthalpy of combustion products at the temperature at the pass
H P= 21171.8 kJ/kg.
Heat flux taken up by water vapor in radiant tubes:
where H n is the enthalpy of combustion products at flue gas temperature at the pass, kJ/kg;
η t - coefficient useful action furnaces; it is recommended to take it equal to 0.95 - 0.98;
Heat flux taken up by water vapor in convection pipes:
The enthalpy of water vapor at the entrance to the radiant section will be:
kJ/kg.
We accept the value of pressure losses in the convection chamber ∆ P to= 0.1 MPa, then:
P to = P - P to ,
P to= 1.2 - 0.1 = 1.1 MPa.
Temperature of water vapor inlet to the radiant section t to= 294 °C, then the average temperature of the outer surface of the radiant tubes will be:
where Δt- the difference between the temperature of the outer surface of the radiant pipes and the temperature of the water vapor (raw material) heated in the pipes; Δt= 20 - 60 °С;
TO.
Maximum design combustion temperature:
where t o- reduced temperature of the initial mixture of fuel and air; taken equal to the temperature of the air supplied for combustion;
THX.- specific heat capacity of combustion products at temperature t P;
°C.
At tmax = 1772.8 °С and t n \u003d 800 ° C heat density of an absolutely black surface qs for various temperatures of the outer surface of radiant tubes has the following values:
Θ, °С 200 400 600
qs, W/m2 1.50 ∙ 10 5 1.30 ∙ 10 5 0.70 ∙ 10 5
We build an auxiliary chart (Fig. 2) see Attachment, according to which we find the heat density at Θ = 527 °С: qs\u003d 0.95 ∙ 10 5 W / m 2.
We calculate the total heat flux introduced into the furnace:
Preliminary value of the area equivalent to a completely black surface:
m 2.
We accept the degree of masonry screening Ψ = 0.45 and for α = 1.25 we find that
Hs /H l = 0,73.
The value of the equivalent flat surface:
m 2.
We accept a single-row placement of pipes and a step between them:
S = 2d n= 2 ∙ 0.152 = 0.304 m. For these values, the form factor To = 0,87.
The value of the shielded masonry surface:
m 2.
Heating surface of radiant tubes:
m 2.
We choose the BB2 oven, its parameters:
radiation chamber surface, m 2 180
convection chamber surface, m 2 180
working length of the furnace, m 9
radiation chamber width, m 1.2
version b
fuel combustion method flameless
radiation chamber pipe diameter, mm 152×6
convection chamber pipe diameter, mm 114×6
Number of pipes in the radiation chamber:
where d n - outside diameter pipes in the radiation chamber, m;
l floor - the useful length of the radiant pipes, washed by the flow of flue gases, m,
l floor = 9 - 0.42 = 8.2 m,
.
Thermal stress of the surface of radiant tubes:
W / m 2.
Determine the number of convection chamber pipes:
We arrange them in a checkerboard pattern, 3 in one horizontal row. Step between pipes S = 1.7 d h = 0.19 m.
The average temperature difference is determined by the formula:
°C.
Heat transfer coefficient in the convection chamber:
W / (m 2 ∙ K).
The heat stress of the surface of convection pipes is determined by the formula:
W / m 2.
2.4 Hydraulic calculation of the furnace coil
The hydraulic calculation of the furnace coil consists in determining the pressure loss of water vapor in radiant and convection pipes.
where G
ρ to v.p. - the density of water vapor at an average temperature and pressure in the convection chamber, kg / m 3;
d k – internal diameter of convection pipes, m;
z k is the number of flows in the convection chamber,
m/s.
ν k \u003d 3.311 ∙ 10 -6 m 2 / s.
The value of the Reynolds criterion:
m.
Friction pressure loss:
Pa = 14.4 kPa.
Pa = 20.2 kPa.
where Σ ζ to
- number of turns.
Total pressure loss:
2.5 Calculation of water vapor pressure loss in the radiation chamber
Average steam velocity:
where G is the flow rate of water vapor superheated in the furnace, kg/s;
ρ r v.p. - the density of water vapor at an average temperature and pressure in the convection chamber, kg / m 3;
dр – internal diameter of convection pipes, m;
z p is the number of flows in the clnvection chamber,
m/s.
Kinematic viscosity of water vapor at average temperature and pressure in the convection chamber ν p \u003d 8.59 ∙ 10 -6 m 2 / s.
The value of the Reynolds criterion:
Total length of pipes in a straight section:
m.
Hydraulic friction coefficient:
Friction pressure loss:
Pa = 15.1 kPa.
Pressure loss to overcome local resistance:
Pa = 11.3 kPa,
where Σ ζ p\u003d 0.35 - resistance coefficient when turning by 180 ºС,
- number of turns.
Total pressure loss:
The calculations performed showed that the selected furnace will provide the process of superheating water vapor in a given mode.
3. Calculation of the waste heat boiler
Find the average flue gas temperature:
where t 1 - flue gas temperature at the inlet,
t 2 – outlet flue gas temperature, °С;
°C (538 K).
Flue gas mass flow:
where B - fuel consumption, kg / s;
For flue gases, specific enthalpies are determined based on the data in Table. 3 and fig. 1 according to the formula:
Enthalpies of coolants Table 4
Heat flux transmitted by flue gases:
where H 1 and H 2 - enthalpy of flue gases at the temperature of the inlet and outlet of the KU, respectively, formed during the combustion of 1 kg of fuel, kJ/kg;
B - fuel consumption, kg/s;
h 1 and h 2 - specific enthalpies of flue gases, kJ / kg,
Heat flux perceived by water, W:
where η ku - coefficient of heat utilization in CU; η ku = 0.97;
G n - steam capacity, kg/s;
h k vp - enthalpy of saturated water vapor at outlet temperature, kJ/kg;
h n in - feed water enthalpy, kJ/kg,
The amount of water vapor received in the KU is determined by the formula:
kg/s.
Heat flux taken up by water in the heating zone:
where h k in - specific enthalpy of water at the evaporation temperature, kJ / kg;
Heat flux transferred by flue gases to water in the heating zone (useful heat):
where h x is the specific enthalpy of flue gases at temperature t x , from here:
kJ/kg.
The value of the enthalpy of combustion of 1 kg of fuel:
According to fig. 1 flue temperature corresponding to the value H x = 5700.45 kJ/kg:
t x = 270 °С.
Average temperature difference in the heating zone:
°C.
270 flue gases 210 Taking into account the counterflow index:
where To f is the heat transfer coefficient;
m 2.
Average temperature difference in the evaporation zone:
°C.
320 flue gases 270 Taking into account the counterflow index:
187 water vapor 187
Heat exchange surface area in the heating zone:
where To f is the heat transfer coefficient;
m 2.
Total heat exchange surface area:
F = F n + F u ,
F\u003d 22.6 + 80 \u003d 102.6 m 2.
In accordance with GOST 14248-79, we select a standard evaporator with a vapor space with the following characteristics:
casing diameter, mm 1600
number of tube bundles 1
number of pipes in one bundle 362
heat exchange surface, m 2 170
sectional area of one stroke
through pipes, m 2 0.055
4. Heat balance of the air heater
Atmospheric air with temperature t ° in-x enters the apparatus, where it is heated to a temperature t x in-x due to the heat of flue gases.
Air consumption, kg / s is determined based on required amount fuel:
where AT- fuel consumption, kg/s;
L- actual air consumption for combustion of 1 kg of fuel, kg/kg,
Flue gases, giving off their heat, are cooled from t dg3 = t dg2 before t dg4 .
=
where H3 and H4- flue gas enthalpies at temperatures t dg3 and t dg4 respectively, kJ/kg,
Heat flux perceived by air, W:
where with in-x- average specific heat capacity of air, kJ/(kg K);
0.97 - air heater efficiency,
Final air temperature ( t x in-x) is determined from the heat balance equation:
TO.
5. Heat balance of KTAN
After the air heater, flue gases enter the contact apparatus with an active nozzle (KTAN), where their temperature decreases from t dg5 = t dg4 up to temperature t dg6= 60 °С.
The flue gas heat is removed by two separate water flows. One stream comes into direct contact with flue gases, and the other exchanges heat with them through the coil wall.
Heat flow given off by flue gases, W:
where H5 and H6- flue gas enthalpies at temperature t dg5 and t dg6 respectively, kJ/kg,
The amount of cooling water (total), kg/s, is determined from the heat balance equation:
where η - KTAN efficiency, η=0.9,
kg/s.
Heat flux perceived by cooling water, W:
where G water- cooling water consumption, kg/s:
with water- specific heat capacity of water, 4.19 kJ/(kg K);
t n water and t to water- water temperature at the inlet and outlet of the KTAN, respectively,
6. Calculation of the efficiency of the heat recovery plant
When determining the value of the efficiency of the synthesized system ( η mu) the traditional approach is used.
The calculation of the efficiency of the heat recovery plant is carried out according to the formula:
7. Exergy assessment of the "furnace - waste heat boiler" system
The exergetic method of analysis of energy technological systems allows the most objective and qualitative assessment of energy losses, which are not detected in any way during a conventional assessment using the first law of thermodynamics. In the case under consideration, the exergy efficiency is used as an evaluation criterion, which is defined as the ratio of the exergy removed to the exergy supplied to the system:
where E sub- fuel exergy, MJ/kg;
E resp.- exergy taken up by the flow of water vapor in the furnace and waste heat boiler.
In the case of gaseous fuel, the supplied exergy is the sum of the exergy of the fuel ( E sub1) and air exergy ( E sub2):
where N n and But- air enthalpies at the furnace inlet temperature and ambient temperature, respectively, kJ/kg;
That- 298 K (25 °С);
∆S- change in air entropy, kJ/(kg K).
In most cases, the value of air exergy can be neglected, that is:
The allotted exergy for the system under consideration is the sum of the exergy taken up by water vapor in the furnace ( E resp1), and the exergy taken up by water vapor in the CH ( E resp2).
For steam flow heated in a furnace:
where G- steam consumption in the furnace, kg/s;
H vp1 and H vp2- enthalpies of water vapor at the inlet and outlet of the furnace, respectively, kJ/kg;
ΔS vp- change in entropy of water vapor, kJ/(kg K).
For the flow of water vapor obtained in the HV:
where G n- steam consumption in CU, kg/s;
h to ch- enthalpy of saturated water vapor at the outlet of the KU, kJ/kg;
h n in- enthalpy of feed water at the inlet to the KU, kJ/kg.
E resp. = E otv1 + E otv2 ,
E resp.\u003d 1965.8 + 296.3 \u003d 2262.1 J / kg.
Conclusion
Having carried out the calculation for the proposed installation (recovery of the heat of the waste gases of the process furnace), we can conclude that for a given fuel composition, furnace productivity in terms of water vapor, and other indicators, the efficiency of the synthesized system is high, so the installation is effective; this was also shown by the exergy assessment of the system "furnace - waste heat boiler", however, in terms of energy costs, the installation leaves much to be desired and needs to be improved.
List of used literature
1. Haraz D .And. Ways of using secondary energy resources in chemical industries / D. I. Kharaz, B. I. Psakhis. - M.: Chemistry, 1984. - 224 p.
2. Scoblo A . And. Skoblo A.I., Tregubova I.A., Yu.K., Molokanov. - 2nd ed., revised. and additional – M.: Chemistry, 1982. – 584 p.
3. Pavlov K .F. Examples and tasks in the course of processes and apparatuses of chemical technology: Proc. Manual for universities / K. F. Pavlov, P. G. Romankov, A. A. Noskov; Ed. P. G. Romankova. - 10th ed., revised. and additional - L.: Chemistry, 1987. - 576 p.
Application
When constructing a furnace, ideally, one would like to have such a design that would automatically give as much air as needed for combustion. At first glance, this can be done with chimney. Indeed, the more intensely the firewood burns, the more hot flue gases should be, the greater the thrust (carburetor model) should be. But it's not. The draft does not depend at all on the amount of hot flue gases generated. Draft is the pressure drop in the pipe from the pipe head to the firebox. It is determined by the height of the pipe and the temperature of the flue gases, or rather, their density.
Thrust is determined by the formula:
F \u003d A (p in - p d) h
where F is the thrust, A is the coefficient, p in is the density of the outside air, p d is the flue gas density, h is the height of the pipe
Flue gas density is calculated by the formula:
p d \u003d p in (273 + t in) / (273 + t d)
where t in and t d - temperature in degrees Celsius of the outside atmospheric air outside the pipe and flue gases in the pipe.
Flue gas velocity in the pipe ( volume flow, that is, the suction capacity of the pipe) G does not depend on the height of the pipe at all and is determined by the temperature difference between the flue gases and the outside air, as well as the area cross section chimney. A number of practical conclusions follow from this.
Firstly, the chimneys are made high not at all in order to increase the air flow through the firebox, but only to increase the draft (that is, the pressure drop in the pipe). This is very important to prevent tipping of the draft (furnace smoking) in case of wind pressure (the thrust value must always exceed the possible wind pressure).
Secondly, it is convenient to regulate the air flow with the help of devices that change the area of \u200b\u200bthe free section of the pipe, that is, with the help of valves. With an increase in the cross-sectional area of the chimney channel, for example, by a factor of two, one can expect an approximately twofold increase in the volumetric air flow through the firebox.
Let's make it simple and good example. We have two identical ovens. We combine them into one. We get a double-sized stove with double the amount of burning wood, with double the air flow and the cross-sectional area of \u200b\u200bthe pipe. Or (which is the same thing), if more and more firewood flares up in the firebox, then it is necessary to open the valves on the pipe more and more.
Thirdly If the stove burns normally in the steady state, and we additionally let a stream of cold air into the firebox past the burning wood into the chimney, then the flue gases will immediately cool down, and the air flow through the stove will decrease. At the same time, burning firewood will begin to fade. That is, we do not seem to directly affect the firewood and direct the additional flow past the firewood, but it turns out that the pipe can pass less flue gases than before, when this additional air flow was absent. The pipe itself will reduce the air flow to the firewood that was previously, and besides, it will not let in an additional flow of cold air. In other words, the chimney will be blocked.
That is why cold air leaks through cracks in the chimneys, excessive air flows in the firebox, and indeed any heat loss in the chimney leading to a decrease in the temperature of the flue gases are so harmful.
Fourth, the greater the coefficient of gas-dynamic resistance of the chimney, the lower the air flow. That is, it is desirable to make the walls of the chimney as smooth as possible, without turbulences and without turns.
Fifth, the lower the temperature of the flue gases, the more sharply the air flow changes with fluctuations in the temperature of the flue gases, which explains the situation of the instability of the pipe when the furnace is ignited.
At sixth, at high flue gas temperatures, the air flow rate is independent of the flue gas temperature. That is, with a strong heating of the furnace, the air flow ceases to increase and begins to depend only on the cross section of the pipe.
Instability issues arise not only when analyzing the thermal characteristics of a pipe, but also when considering the dynamics of gas flows in a pipe. Indeed, the chimney is a well filled with light flue gas. If this light flue gas does not rise very quickly, then there is a possibility that the heavy outside air can simply sink into the light gas and create a falling downward flow in the pipe. This situation is especially likely when the walls of the chimney are cold, that is, during the ignition of the furnace.
Rice. 1. Scheme of the movement of gases in a cold chimney: 1 - firebox; 2 - air supply through the blower; 3-chimney; 4 - valve; 5 - chimney tooth; 6-flue gases; 7-failing cold air; 8 - air flow causing thrust tipping.
a) smooth open vertical pipe
b) a pipe with a valve and a tooth
c) pipe with top valve
Solid arrows show the directions of movement of light hot flue gases. Dashed arrows show the directions of downward flows of cold heavy air from the atmosphere.
On the rice. 1a a furnace is schematically shown, into which air 2 is supplied and flue gases 6 are removed through the chimney. even the firebox. This falling flow can replace the “regular” air flow through blower 2. Even if the stove is locked with all the doors and all air intake dampers are closed, the stove can still burn due to the air coming from above. By the way, this is what often happens when coal burns out with the oven doors closed. A complete tipping of the draft may even occur: air will enter from above through the pipe, and flue gases will exit through the door.
In reality, on the inner wall of the chimney there are always bumps, growths, roughness, upon collision with which the flue gases and the oncoming downward cold air flows swirl and mix with each other. At the same time, the cold downward air flow is pushed out or, heating up, begins to rise upwards mixed with hot gases.
The effect of turning downward flows of cold air upwards is enhanced in the presence of partially open dampers, as well as the so-called tooth, which is widely used in the technology of manufacturing fireplaces ( rice. 1b). The tooth prevents the flow of cold air from the chimney into the fireplace space and thereby prevents the fireplace from smoking.
Downdrafts of air in the pipe are especially dangerous in foggy weather: flue gases are not able to evaporate the smallest droplets of water, they cool down, the thrust decreases and may even tip over. At the same time, the stove smokes heavily, does not flare up.
For the same reason, stoves with damp chimneys smoke a lot. Top gate valves are particularly effective in preventing downflows ( rice. 1v), adjustable depending on the speed of the flue gases in the chimney. However, the operation of such valves is inconvenient.
Rice. Fig. 2. Dependence of the excess air coefficient a on the time of heating the furnace (solid curve). The dotted curve is the required air consumption G consumption for the complete oxidation of the combustion products of firewood (including soot and volatile substances) in flue gases (in relative units). The dashed-dotted curve is the actual air consumption G of the pipe provided by the pipe draft (in relative units). The excess air coefficient is the quotient of G pipe separation per G flow
A stable and sufficiently strong draft occurs only after the walls of the chimney have warmed up, which takes a long time, so there is always not enough air at the beginning of the heating. In this case, the excess air coefficient is less than unity, and the furnace smokes ( rice. 2). And vice versa: at the end of the heating, the chimney remains hot, the draft remains for a long time, although the firewood has almost burned out (the excess air coefficient is more than one). Metal furnaces with metal insulated chimneys reach the regime faster due to their low heat capacity compared to brick chimneys.
The analysis of the processes in the chimney can be continued, but it is already clear that no matter how good the stove itself is, all its advantages can be reduced to zero by a bad chimney. Of course, ideally, the chimney would need to be replaced modern system forced extraction of flue gases by means of an electric fan with adjustable flow rate and with pre-condensation of moisture from flue gases. Such a system, among other things, could clean the flue gases from soot, carbon monoxide and other harmful impurities, as well as cool the discharged flue gases and provide heat recovery.
But all this is in the distant future. For a summer resident and gardener, a chimney can sometimes become much more expensive than the stove itself, especially in the case of heating a multi-level house. Sauna chimneys are usually simpler and shorter, but the heat output level of the stove can be very high. Such pipes, as a rule, are very hot along the entire length, sparks and ash often fly out of them, but condensation and soot are insignificant.
If for now you plan to use the sauna building only as a bathhouse, then the pipe can also be made uninsulated. If you also think about the bathhouse as a place of possible stay (temporary residence, overnight stays), especially in winter, then it is more expedient to immediately make the pipe insulated, and qualitatively, “for life”. At the same time, stoves can be changed at least every day, the design can be selected more conveniently and more appropriately, and the pipe will be the same.
At least if the oven is on long burning(smoldering firewood), then pipe insulation is absolutely necessary, because at low power (1 - 5 kW) an uninsulated metal pipe will become completely cold, condensate will flow abundantly, which in the most severe frosts can even freeze and block the pipe with ice. This is especially dangerous in the presence of a spark arresting grid and umbrellas with small passage gaps. Spark arresters are useful for intensive heating in summer and extremely dangerous for weak firewood burning conditions in winter. Due to the possible clogging of pipes with ice, the installation of deflectors and umbrellas on chimneys was banned in 1991 (and on chimneys gas ovens even earlier).
For the same reasons, you should not get carried away with the height of the pipe - the level of thrust is not so important for a non-return sauna stove. If it smokes, you can always quickly ventilate the room. But the height above the roof ridge (at least 0.5 m) must be observed to prevent tipping of the thrust during gusts of wind. On flat roofs, the pipe should protrude above the snow cover. In any case, it is better to have a lower pipe, but warmer (than higher, but colder). Tall chimneys are always cold and dangerous in winter.
Cold chimneys have a lot of disadvantages. At the same time, non-insulated, but not very long pipes on metal stoves warm up quickly during kindling (much faster than brick pipes), remain hot with vigorous heating, and therefore are very widely used in baths (and not only in baths), especially since they are relatively cheap. Asbestos-cement pipes are not used on metal furnaces, since they are heavy, and also collapse when overheated with fragments flying.
Rice. 3. The simplest designs of metal chimneys: 1 - a metal round chimney; 2 - spark catcher; 3 - a cap to protect the pipe from atmospheric precipitation; 4 - rafters; 5 - roof sheathing; 6 - wooden blocks between rafters (or beams) to design a fire opening (cutting) in the roof or ceiling (if necessary); 7 - roof ridge; eight - soft roof(roofing material, hydrostekloizol, soft tiles, corrugated cardboard-bitumen sheets, etc.); 9 - metal sheet for roofing and covering the opening (it is allowed to use flat sheet aceida - asbestos-cement electrical insulating board); 10 - metal drainage pad; 11 - asbestos sealing of the gap (joint); 12 - metal cap-otter; 13 - ceiling beams(with filling the space with insulation); 14 - ceiling lining; 15 - attic floor (if necessary); 16 - metal sheet of ceiling cutting; 17 - metal reinforcing corners; 18 - metal cover of the ceiling cut (if necessary); 19 - non-combustible heat-resistant insulation (expanded clay, sand, perlite, mineral wool); 20 - protective pad (metal sheet over a layer of asbestos cardboard 8 mm thick); 21 - metal pipe screen.
a) non-thermal insulated pipe;
b) a heat-insulated shielded pipe with a heat transfer resistance of at least 0.3 m 2 -deg / W (which is equivalent to a brick thickness of 130 mm or a mineral wool insulation thickness of 20 mm).
On the rice. 3 shows typical wiring diagrams for non-insulated metal pipes. The pipe itself should be purchased from stainless steel with a thickness of at least 0.7 mm. The most popular diameter of the Russian pipe is 120 mm, the Finnish one is 115 mm.
According to GOST 9817-95, the cross-sectional area of a multi-turn chimney should be at least 8 cm 2 per 1 kW of rated heat output released in the furnace when wood is burned. This power should not be confused with the thermal power of a heat-intensive furnace, released from the outer brick surface of the furnace into the room according to SNiP 2.04.05-91. This is one of our many misunderstandings. normative documents. Since heat-intensive furnaces are usually heated only 2-3 hours a day, the power in the furnace is about ten times greater than the power of heat release from the surface of a brick oven.
Next time we will talk about the features of the installation of chimneys.