CC BY
853 Wschodnioeuropejskie Czasopismo Naukowe (East European Scientific Journal)#2(30), 2018 sfesLH
M. Bezrodny, N. Prytula, T. Misiura National Technical University of Ukraine «Igor Sikorsky Kyiv Polytechnic Institute»
THERMODYNAMIC EFFICIENCY OF COMBINED HEAT PUMP SYSTEM OF HEATING AND VENTILATION WITH USE OF HEAT OF VENTILATING EMISSIONS AND WASTEWATER
The efficiency of the combined heat pump system of heating and ventilation is analyzed with the use of the heat of the ventilation and preheated at the expense of the heat of the sewage of the atmospheric air. A theoretical model of this system is developed and a numerical analysis of its thermodynamic efficiency is performed. It is determined that the use of this system allows to reduce the total specific costs of external energy for heating and ventilation in comparison with the system without using the heat of waste water. Graphic dependencies are obtained with the image of optimal operating conditions of the heat pump system.
Keywords: heat pump, combined heating and ventilation system, waste water, total specific costs of external energy.
1. Introduction
Since natural resources are decreasing each year, and the demand for energy, by contrast, does not cease to grow, the problem of their efficient and economical use is relevant worldwide. Therefore, the search for alternative energy sources that can reduce fuel consumption in conventional plants has a long-term perspective.
In non-traditional energy, heat pumps are the most powerful low-power alternate energy sources. Heat pumps using atmospheric air as a lower source have become widely used due to low investment compared to other sources of energy, unlimited availability and accessibility of heat sources. However, their significant drawback is the loss of power and efficiency with lower air temperatures [1].
For many regions, the temperature potential of atmospheric air is not sufficient for the efficient use of the heat pump system for heating purposes during the cold season. At the same time, some manufacturers offer heat pump equipment, which allows normal operation at ambient temperatures to -20°C. However, in these conditions, the transformation factors are usually below the values adopted for an energy efficient system. Therefore, in order to ensure the operation of the heat pump throughout the heating period, it is necessary to increase the air temperature at the inlet to the evaporator at low ambient air temperature.
There are different approaches to solving this problem. The heat pump circuits for water heating and ventilation with the utilization of the heat of ventilation air is one of these approaches [2, 3]. However, in cottages, residential and public buildings, in which the thermal power of the ventilation system is insignificant compared to the heat consumption for heating, the heat pump is not able to provide a load on heating and ventilation due to only the heat of ventilation emissions. Therefore, there is a need for an additional source of heat.
As such heat sources can be used wastewater. As a result of comparatively constant consumption of heat for hot water supply during the year, significant amounts of conventionally clean wastewater of buildings with a temperature of about 32^ are discharged into the sewage system. Therefore, they can be used as an additional stable and conditionally free source of heat.
The authors proposed and analyzed the principle scheme of the combined heat pump system of heating and ventilation using the heat of ventilation and preheated by the heat of sewage of atmospheric air. As a criterion for thermodynamic efficiency, the value of the total specific energy consumption of external energy is chosen, which represents the amount of energy consumed per unit of heat produced to meet the heating and ventilation needs.
2. Thermodynamic analysis of the combined heat pump system of heating and ventilation
In fig. 1 is shown a schematic diagram of the heat pump system of heating and ventilation using the heat of ventilation and preheated by the heat of waste water of atmospheric air. The principle of operation of the scheme: atmospheric air (the main source of heat) with a temperature to and mass flow of Gat preheated in the air-water heat exchanger due to the heat of wastewater Qw.w to a temperature ti. Heated atmospheric air enters the mixing chamber, where it is mixed with ventilation emissions with temperature ti and mass flow Gvent. The resulting mixture of air (ti, Gmix) after the mixing chamber is directed with the help of a fan to the evaporator of the HP, where the air is cooled and at the outlet has a temperature te. To compensate heat consumption for heating and ventilation, the heat flow from the compressor of the HP Qheat+vent with the temperature of the coolant tc at the entrance to the heating and ventilation system is used.
Fig. 1. Principal scheme of the combined heat pump system of low-temperature water heating and ventilation using the heat of the ventilation and heated by the waste water of atmospheric air: HP - heat pump; CHP -condenser of the HP; EHP - evaporator of the HP; C - compressor; F - fan; H - air heater; MC - mixing
chamber; HE - heat exchanger on sewage.
The thermal load of the evaporator of the HP in
The efficiency of the combined heat pump system of low-temperature water heating and ventilation, taking into account the energy consumption for the drive of the compressor of the HP and the fan, which pumps the air mixture through the evaporator of the HP, can be characterized by the amount of specific energy consumption for heating and ventilation, which is the ratio of spent external energy to the amount of heat received to meet the needs of heating and ventilation
_ A+Lf
l
heat+vent
Q
(1)
heat+vent
where Lc, Lf - power of the compressor drive of the HP and the fan, kW; Qheat+vent - the heat flow, brought to the room to meet the needs of heating and ventilation, kW.
The work of the compressor drive of the HP in the general case is defined as
L --Q
1)
(2)
where Qev - thermal load of the evaporator of the HP, kW; 9 - coefficient of transformation of heat of the HP.
this case is determined by the formula
Qev - GmixCp(ti )
(3)
where Gmix = Gat + Gvent - total mass flow rate of atmospheric Gat and vent Gvent air to evaporator of the HP, kg / s; cp - isobar heat capacity, kJ / (kg °C); ti, te -the temperature of the air at the inlet and outlet of the evaporator of the HP, respectively, °C.
The energy consumption of the fan drive can be determined by the equation
Lf
GL
» 1 -Ap—,
(4)
where pa - air density, kg / m3; Ap - pressure loss in evaporator of the HP, kPa;
n = nf ndr; nf and ndr - the efficiency of the fan and its drive, respectively. It is accepted that in the optimal mode of operation of the fan nf = 0,8, and the efficiency of the drive n<t = 0,95.
The heat flow, brought to the premises to compensate for the loss of heat for heating and ventilation, is determined by the ratio
Q
heat+vent
: Qheat + Qvent - Qvent(1 + 1 / ™) - Gvent(t ~ tQ )(1 + 1 / m) , (5)
This value should be set on the basis of preliminary where m is the ratio of the flow of heat for calculations of heat consumption for heating and ventilation to the flow of heat for heating ventilation.
m
Qvent/ Q
heat *
(6) The thermal equilibrium of the HP allows to
determine the temperature dependence of the mixture
of ventilation and heated air at the outlet of the evaporator of the HP te from the coefficients k and n.
The coefficient k is the ratio of the flow of heat of wastewater Qw.w to the flow of heat for heating Qheat depending on the ambient temperature and is defined as
O t _1r
k = Oww = kr. = k . B (7)
Oheat ti _ t0
where kr - ratio of heat fluxes at the rated ambient temperature (in these calculations, this temperature is -20 °C); ti - the temperature of the air indoors is 20 °C; to is the temperature of the ambient air.
Coefficient n is the ratio of the mass flow of atmospheric air Gat to the total air flow Gmix through the evaporator of the HP and has the form
n
■GJ Gm
(8)
Knowing the coefficients m and k, one can determine the proportion of air heated in the heat exchanger on waste water. According to Fig. 1 heat balance of the heat exchanger can be written as follows
Gat ¿p (t - t0 ) = -Qent . (9)
m
If we divide the left and right sides of equation (9) into Gmix and plot the Qvent values, we obtain the following relation
Gat _ k Gvent¿p (ti - to )
Gsun mGmix ¿p (ti - to )
. (10)
Taking into account the equation (8), after a series of mathematical transformations, we obtain an expression for the dependence of the proportion of air heated in the heat exchanger on waste water from the coefficients m and k (neglecting the change in the specific heat)
-
n
m + -
(11)
Heat transfer transformation coefficient of the HP is defined as
9 = 9tHHP ,
(12)
where ^hp is a coefficient that takes into account the real processes carried out by the working body in the HP, which, according to a number of sources, can vary in the range of 0.6 ... 0.8 (in this case ^hp = 0.6) [4, 5]; ^x - the theoretical coefficient of transformation of a HP.
The heat transfer transformation coefficient of the ideal Carnot cycle ^t, taking into account the thermal non-convergences in the evaporator and the condenser, is determined by the ratio
9t =
T.
HP
T hp _ Y HP
273 +1 + At
tc - te + Atc + Ate
(13)
HP
rp
where 1 e - absolute temperature of evaporation
rp HP
of the refrigerant in the evaporator of a HP, K; 1 c -
absolute temperature of condensation of the refrigerant in the condenser of a HP, K; te - temperature of the exhaust air at the outlet of the evaporator of a HP, °C; tc - water temperature at the outlet of the condenser of a HP, °C; Ate - temperature difference between flows of exhaust air and the refrigerant at the outlet of the evaporator of a heat pump, °C; Atc - temperature difference between the flows of the refrigerant and the heat-carrier in the heating and ventilation system at the outlet of the condenser of a HP, °C. The literature gives numerical values of temperature differences in a condenser and evaporator for air-to-air heat pumps. According to [4], for a HP condenser, Atc = 5 °C can be taken, Ate = 10 °C for the evaporator.
The temperature of the coolant supplied to the heating system, depending on the temperature of the outside air, is determined by the ratio [6]
= ^+( th-11)
where th
'( tj-10) " ( ti-1o)
(1+n)
(14)
th - the rated temperature of the heating
fluid in the heating system at the rated temperature of the atmospheric air; n = 0 for low temperature heating systems. The rated temperature of the coolant in the
heating system is taken th = 45 °C.
Taking into account equations (2) - (5), the ratio (1) for determining the total specific energy expenditure for combined heat and power supply heating and ventilation will be as follows
GmixCp (ti - te ) / (9 - 1) + GmiX Ap / Pa^
heat+vent
Gvent¿p (hi - )(1 + 1/ m)
(15)
1
To obtain the final formula for lheat+vent, it is necessary to determine the temperature of the air at the outlet of the evaporator of the HP te.
An analytical expression for determining the temperature of the exhaust air at the outlet of the
evaporator te can be obtained from the heat balance of the heat pump unit
Qv + L = Oc, (16)
where Qc - heat flow, obtained from the condenser of the HP, which is equal to the heat flow for heating and ventilation.
Similarly to equation (5) we define Qc as
Q = Qheat + Qvent = Qheat (1 + . (17)
Taking into account equations (2), (3) and (17) the expression (16) takes the form
GmlxCp (t - te = Qheat (1 + m) . (18)
p m -1
The complex of values can be obtained from the energy balance of the system as a whole. According to Fig. 1, the energy balance of the combined heat pump heat and ventilation systems can be written as follows
L + Gent^o + L + GCpti = Qheat + GmixVe . (20)
Then, taking into account (2) - (4) and (8), from equation (20) we obtain the following relation
= (1 - n)tfj + mf. + ii^e! + A1 - ^, (21) GmixS m-1 n
By solving the relation (18) with respect to te, an analytical expression can be written to determine the temperature of the exhaust air mixture at the outlet of the evaporator of the HP
Q
heat
G . c
mix p
(1 + m) 1. (19) 9
where,
Ä--
¿P /(PaCp )
°C is the ratio of the
given values.
Substituting equation (21) in expression (19) and following a series of mathematical transformations, we obtain the finite correlation for determining the temperature of the air mixture at the outlet of the evaporator of the HP
t
t0 (1 - n)(9 -1)(1 + m) +1 • -1)(1 + m) + Ä(9 -1)(1 + m)/n ~ t [9 - (1 + m)]
te _ .
9 m
(22)
Taking into account (22), after a series of the combined heat pump system of heating and mathematical transformations, the equation (15) for ventilation will have the form estimating the total specific energy consumption of
1
1
heat+vent
(1 - n) (1/m +1)
1
t -1 A
1
9 -1 ti - to n (ti - to)
(23)
3. Numerical analysis of the system
Numerical analysis using relations of (6), (8), (11), (12), (14), (22), (23) by the method of successive approximations allows to estimate the influence of changes in the temperature of the environment, the ratio of the flow of heat to ventilation to heat flux for heating and the ratio of the flow of heat to wastewater to the heat flow for heating as the system parameters (the temperature of the mixture of air at the exit from the evaporator of the HP (Figure 2), the coefficient of transformation of the HP (Figure 3)), and on the thermodynamic efficiency of HP application of water heating and ventilation using heat mixture and heated ventilation air (Fig. 4). When choosing a value m this circumstance is taken into account. It is known that the relative cost of heat for the ventilation of residential and office space is about 20 ... 100% of the cost of heat for heating. In this connection, the following values were taken in this analysis m = 0.2; 0.5; 0.8 With cost analysis of heat for hot water supply
and heating coefficient kr takes the following values: kr = 0.2; 0.3; 0.4 It was also conducted to compare the results provided when using only heat vent emissions, i.e. when n = 0. A set value ratio A, which is set for the average real value of HP evaporator aerodynamic resistance as the convective heat exchanger is 0.5.
In fig. 2, a-c are shown the graphical dependences of the temperature of the air mixture at the outlet of the evaporator of the HP on the temperature of the atmospheric air, from which it is evident that the temperature te increases with an increase in the coefficients m and kr. It should also be noted that some temperatures te even exceed the corresponding values of ambient temperatures, which indicates an improvement in the conditions of the HP. The temperature of the mixture of air te in the investigated scheme is higher than in the original. But with increasing flow of heat for ventilation (at m> 1) this effect is reduced.
te = f{t0), m = 0.5
20 10
0 -10
-20 -30 -40 -50 -60 -70
l 2 3 -1- J
/ / _-i- * ^ *
I-J z.
__ _ - - 1 1 ■ " A "...__-p
" ' "" 1 a jh **
-20
-15
-10
-5
10
b)
*e = fit
o), m = 0.8
20 10
0 -10 -20 -30 -40 -50 -60 -70
-20
-15
-10
10
15
1 1 j 4
! /
/ J - m***
7" —' ' M -1
mm- * —
...............
15
c)
Fig. 2. Dependence of the temperature of the air mixture at the outlet of the evaporator of the HP on the ambient temperature: a), b), c) - m = 0.2; 0.5; 0.8 respectively; 1-4 - kr = 0.2; 0.3; 0.4; n = 0.
In fig. 3, a-c are shown the graphical depend- the transformation coefficient 9 increases with in-ences of the transformation coefficient of the HP on creasing coefficients m and kr and the ambient temper-the ambient temperature, from which it is evident that
ature. The value of the transformation coefficient indicates that the heat utilization of the heated air and the ventilation emissions in the heating and ventilation systems are sufficient. It can also be seen that the transformation coefficient in the investigated scheme
is larger compared to the use of only ventilation emissions. The difference between these two cases increases with an increase in the temperature of the environment. But with increasing m, this advantage is reduced.
a)
12 11 10
9 8 7 6 5 4 3 2 1
<P
<p= f(t0), m=0.5
1 2 3 4
j / j s'y
/ / '■Y
/ / / * s . / y /
j /
/ **I**
/ _ j „
_ — — F ■■......"
---Jj-S trrm ..........
-20
-15
-10
-5
10
15 l0
b)
Fig. 3. Dependence of the coefficient of transformation of the HP on ambient temperature: a), b), c) - m = 0.2; 0.5;
0.8 respectively; 1-4 - kr = 0.2; 0.3; 0.4; n = 0.
On the basis of the numerical analysis of the relation (23) taking into account the obtained values of such parameters as the coefficients n and m, the temperature of the mixture of air at the outlet of the evaporator of the HP and the heat transfer transformation coefficient of the HP, the graphic dependences of the specific energy expenditure on the ambient temperature are plotted (Fig. 4, a-c). It can be seen that specific costs decrease with increasing ambient temperature and with the increase of the coefficients m and k'. The advantage of using the
scheme under study compared with the original decreases with an increase in the value of m, as can be seen in Fig. 4, c. At m > 1 this advantage disappears. This indicates that with the further growth of the heat flow to ventilation, the use of warmth of heated air is meaningless, and general needs of heat can be satisfied only at the expense of ventilation emissions. The presence of a minimum indicates the change in the dominant contribution of the compressor and fan, which occurs only with low heat consumption for heating and ventilation (at high temperatures to).
0.65 0.6 0.55 0.5 0.45 0.4 0.35 0.3 0.25 0.2 0.15 0.1
heat+vent
'heat+vent ~ f{*o)j- m-0.2
i 2 | m
/ / /
J /
/'"•■: / /
^ r r J 1
- - ^ -- * ■ ^ ^ / /
-20
-15
-10
-5
10
15
to
a)
I
0.55 0.6 0.55 0.5 0.45 0.4 0.35 0.3 0.25 0.2 0.15 0.1
heat+vent
heat+vent
= f{t0), m=0.8
1 3 3 4
/ V , ! y
/ / / /
/ / /
.......... / /
..... / *
S S
/fo
-20
-15
-10
c)
10
15 to
Fig. 4. Dependence of the specific energy consumption of external energy on the ambient temperature: a), b), c) - m = 0.2; 0.5; 0.8 respectively; 1-4 - kr = 0.2; 0.3; 0.4; n = 0.
4. Conclusions
1. Combined use of heat of ventilation and heated by the heat of sewage of atmospheric air compared with the use of heat only of ventilation air is more effective. It is also important that this advantage manifests itself to a greater extent at low values of the coefficient m corresponding to the real value of this value for residential and office premises. At m > 1 this advantage disappears.
2. Reduction of specific energy consumption in the system under study at low values of the coefficients m and kr (m = 0.2 and kr = 0.2) reaches 24-50% over the whole range of ambient air temperatures (-20 °C to +10 °C).
3. The efficiency of the heat pump, characterized by the conversion factor (COP), for this combined heat pump system increases throughout the range of temperatures of the outside air, but most significantly in the zone of favorable temperatures.
References
1. Gershkovich, V. F. (2009). Features of the design of heating systems of buildings with heat pumps. Kyiv, Ukraine: Ukrainian Academy of Architecture "Energominimum", 60.
2. Bezrodny, M. K., Prytula, N.O. (2011). Energy efficiency of the low-temperature water heating and ventilation combined heat pump system. Naukovi visti NTUU «KPI», 1, 19-25.
3. Bezrodny, M. K., Prytula, N. O. (2012). Energy efficiency of heat pump heating schemes. Kyiv, Ukraine: NTUU "KPI", 208.
4. Morozjuk, T. V. (2006). The theory of chillers and heat pumps. Odessa, Ukraine: Studija « Négociant», 712.
5. Steward, F. R. (1984). Optimum arrangement and use of heat pumps in recovery waste heat. Energy Conversion Mgmt, Vol. 24, 2, 123-129.
6. Shubin, E. P. (1979). The main issues of cities' heating systems designing. Moscow, USSR: Jener-gija, 359.
Index
- tc - temperature of atmospheric air, °C;
- Gat - mass flow of atmospheric air, kg/s;
- Qw.w - heat of wastewater, kW;
- ti - the temperature of the air indoors, °C;
- Gvent mass flow of ventilation emissions,
kg/s;
- Gmix mass flow of resulting mixture of air,
kg/s;
- HP - heat pump;
- te - temperature of outlet air (after the evaporator of the HP), °C;
- Qheat+vent - heat consumption for heating and ventilation, kW;
- tc - temperature of the coolant at the entrance to the heating and ventilation systems, °C;
- Chp - condenser of the HP;
- Ehp - evaporator of the HP;
- C - compressor;
- F - fan;
- H - air heater;
- MC - mixing chamber;
- HE - heat exchanger on sewage;
- lheat+vent - the amount of specific energy consumption for heating and ventilation, kW;
- Lc - power of the compressor drive of the HP,
kW;
- Lf - power of the fan, kW;
- Qev - thermal load of the evaporator of the HP,
kW;
- pa - air density, kg / m3;
- k - ratio of heat fluxes at the rated ambient temperature;
- cp - isobar heat capacity, kJ / (kg °C);
- nf and ndr - the efficiency of the fan and its drive, respectively;
- m - the ratio of the flow of heat for ventilation to the flow of heat for heating;
- n - the ratio of the mass flow of atmospheric air to the total air flow through the evaporator of the HP;
- nHP - a coefficient that takes into account the real processes carried out by the working body in the HP;
- 9t - the theoretical coefficient of transformation of a HP;
- tc - the temperature of the coolant supplied to the heating system;
- th - the rated temperature of the heating fluid
in the heating system at the rated temperature of the atmospheric air;
- Qc - heat flow, obtained from the condenser of the HP, which is equal to the heat flow for heating and ventilation, kW.
Трембус 1рина Вталивна
кандидат техтчних наук, доцент кафедри екологИ та технологирослинних полгмергв, Нацюнальний технгчний утверситет Украти «Кшвський полтехнгчний iнститут 1мет 1горя
Скорського» Соколовська Нша Валеривна астрант кафедри екологИ та технологи рослинних полiмерiв, Нацюнальний технiчний утверситет Украти «Кшвський полiтехнiчний тститут iменi 1горя
Скорського» I.V. Trembus
Ph.D., аssociate professor department of Ecology and technology ofplant polymers, National Technical
University of Ukraine "Igor Sikorsky Kyiv Polytechnic Institute "
N. V. Sokolovska
graduate student of Ecology and technology ofplant polymers, National Technical University of Ukraine "Igor Sikorsky Kyiv Polytechnic Institute"
ДЕЛ1ГШФ1КАЦШ ПШЕНИЧНО1 СОЛОМИ В СИСТЕМ1 СН3СООН - Н2О - Н2О2
DELIGNIFICATION OF WHEAT STRAW IN THE SYSTEM СН3СООН - Н2О - Н2О2
Анотащя. Розглянуто вплив витрат пероксиду водню та тривалосп процесу делгтфжаци пшенично! соломи в середовищi оцтово! кислоти на показники якосп солом'яно! целюлози. Експериментально вста-новлено рацюнальш витрати окисника, що забезпечуе високий вихвд целюлози з низьким вмютом литану. Розраховано показники вибiрковостi вилучення литану для дослщженого окисного способу делптафжацп стебел пшенично! соломи. Запропоновано лптан-вуглеводну дiаграму дел^шфжацп рослинно! сировини. Визначено ряд, в який розташовуються за ефектившстю рiзнi способи дел^шфжацп пшенично! соломи.
Ключовi слова: пшенична солома, окисна делптафжащя, целюлоза, селективнiсть, лiгнiн-вуглеводна дiаграма
Abstract. The influence of hydrogen peroxide and duration of the wheat straw delignification process in acetic acid medium on the quality of pulp with wheat straw quality is considered. Rational consumption of oxidizer is experimentally established that provides high yield of pulp with the low maintenance of a lignin. Indexes of selectivity of extraction of a lignin for the studied oxidizing way of delignification of stalks of wheat straw are calculated. It is offered a lignin - carbohydrate chart of delignification of vegetable raw materials. A row in which various ways of delignification of wheat straw settle down by efficiency is defined.
Key words: wheat straw, oxidizing delignification, pulp, selectivity, lignin -carbohydrate chart
Постановка проблеми. Технолопчш i еколо-ri4Hi проблеми при одержанш целюлози потребу-ють нових техшчних ршень, включаючи вирь шення проблеми сировинно1 бази. Саме ïï вщсутшсть е основним бар'ером для збшьшення обсяпв целюлозно - паперового виробництва. Тому необхвдно здiйснювати пошук нових i розши-рити використання вже ввдомих джерел волокнисто! сировини. До таких джерел вщносять недеревну рослинну сировину (НДРС) i в першу чергу вiдходи сiльського господарства.
Недеривна рослинна сировина привертае увагу вчених, як сировина для хiмiчного перероблення так i для одержання техшчно! целюлози. Основною
перевагою тако! сировини е ïï щорiчна вщновлюва-шсть i бiльш низька вартiсть у порiвняннi з деревиною [1].
У зв'язку з введенням у дш бiльш жорстких вимог до якосп стiчних вод i газових виквдв в атмосферу целюлозно-паперових пiдприемств, а та-кож у зв'язку з урбашзащею промислових тдпри-емств, виникла необхiднiсть впровадження б№ш екологiчних та ресурсозюерыгаючих способiв дель гнiфiкацiï рослинноï сировини [2]. Серед альтерна-тивних технологiй одержання целюлози вченими широко проводяться дослiдження органосольвент-них способiв делiгнiфiкацiï рослинноï сировини. Таю варшня характеризуються бiльшою вибiрко-вою дiею на лiгнiн, що дае можливють збiльшити
