WO2016134731A2 - The ideal liquid compression refrigeration cycle - Google Patents
The ideal liquid compression refrigeration cycle Download PDFInfo
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- WO2016134731A2 WO2016134731A2 PCT/EG2015/000005 EG2015000005W WO2016134731A2 WO 2016134731 A2 WO2016134731 A2 WO 2016134731A2 EG 2015000005 W EG2015000005 W EG 2015000005W WO 2016134731 A2 WO2016134731 A2 WO 2016134731A2
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
- F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
- F25B23/00—Machines, plants or systems, with a single mode of operation not covered by groups F25B1/00 - F25B21/00, e.g. using selective radiation effect
Definitions
- the present invention is directed to the mechanical power engineering for refrigeration and heat pumps.
- Refrigeration cycles transfer thermal energy from a region of low temperature to one of higher temperature
- the reversed Carnot cycle is the perfect model for the refrigeration cycle operating between two fixed temperatures, the most ideal cycle, which has the maximum thermal efficiency, and maximum coefficient of performance, and it serves as a standard against which actual refrigerator cycles can be compared
- reversed Camot cycle consist of 4 process, 2 isentropic process for expansion and compression, and 2 isothermal process for heat rejec on and heat absorption.
- the ideal Vapor compression cycle is a cycle built on the principals of the reversed Carnot, but this cycle is deviate from the rev-Carnot for the following reasons:
- the refrigerant shall enter the compressor at the vapor phase, for the compressor operation.
- Throttling valve is used in expansion process (constant enthalpy process) 3 - The heat rejection and absorption at a constant pressure process, for more practicality.
- the intent of this invention is to prove a new ideal refrigerator cycle (the Liquid compression cycle) which has a coefficient of performance higher than the Vapor compression cycle.
- the coefficient of performance for the ideal Vapor compression cycle is lower than the reversed Carnot cycle due to the deviation of the ideal process of the VCC than the rev-Carnot, this means that the ideal VCC will consume more electric power than the rev. Carnot at the same refrigerant heat load.
- the liquid refrigerant pump in the liquid compression cycle is acting the same function of the compressor in the VCC to solve all the above problems.
- Liquid compression cycle is one of the cycles, that can be applied in the refrigeration applications, this cycle has achieved the performance of the reversed Carnot cycle, unlike the vapor compression cycle, where a clear deviation from the reversed Carnot cycle is appeared in its ideal case, the deviation is occur due to the compression process where the refrigerant has to be compressed to a temperature higher than the condensing temperature, and the constant enthalpy process in the expansion valve, where energy loss has occurred due to the irreversibility, these deviations from Carnot cycle have been solved in the Liquid Compression Cycle (LCC)to achieve a thermal efficiency more than the Vapor Compression Cycle (VCC) efficiency, and we will prove that later.
- LCC Liquid Compression Cycle
- Liquid compression cycle consists of 5 processes, 3 isentropic processes, one isothermal process, and one isobaric process, the cycle (T-H) and (T-S) diagrams are shown in fig-1.
- Liquid compression cycle is working between 3 levels of pressure, the refrigerant enter the pump at state 1 as a saturated liquid and compressed from the condenser pressure to a higher level pressure, then the refrigerant enters the expansion nozzle to reach the evaporator pressure, during this expansion process the refrigerant lose a lot of internal energy as well as the pressure is decreasing, these amount of energy is converted to kinetic energy at state 3, then the refrigerant is absorbing heat during the isothermal process in the evaporator to reach state 4 in a 2 phase region, then the pressure is regained in the diffuser by converting a part of the kinetic energy again to enthalpy, the refrigerant is isentropic compressed to the condenser pressure at state 5, then the heat is rejected to the ambient at constant pressure to enter the pump again at state 1, g-4 is showing a schema c diagram for the cycle main components. • Example
- the following example is showing how the Liquid compression cycle has achieved the performance of the reversed Carnot cycle comparing with the Vapor compression cycle at the same levels of condenser and evaporator pressure.
- VCC Vapor Compression cycle
- V s is too small to be considered as a condenser coil velocities or a pump suction velocities, this will lead to slightly increasing the diffuser outlet velocity, and for solving this problem, the process shall be divided in two sections, as shown in the schematic diagram for the diversion - conversion diffuser, (Fig. no. 6)
- Sec on 1 is a diversion diffuser for converting the most amount of kinetic energy to enthalpy energy
- sec on 2 is a conversion nozzle for regaining a small amount of kinetic energy for increasing condenser and pump inlet velocity as shown in the T - H diagram,(Fig. no. 5)
- V 3 A 2 . V 2
- V 2 is representing the pump discharge velocity, so the above value is too small and not practical.
- Sec on 1 is a diversion di user for conver ng small amount of the kine c energy to enthalpy energy till the velocity at the diffuser reach the required value of the nozzle inlet, then the ow will entering sec on 2, where a conversion nozzle is regain most amount of the kinetic energy, till the velocity at the nozzle outlet is reaching the value 3.67 m/s. as shown in the T - H diagram, (Fig. no. 5).
- the higher pressure level is calculated according to the minimum potential work needed for the reversible Liquid compression cycle, in the actual cycle, that pressure shall be increased to overcome the irreversibility in the cycle.
- the COP of the LCC will slightly raised above the reversed Carnot cycle, the limitation for this raise depending on the minimum temperature approach between the refrigerant and the ambient or the cooling medium.
- the LCC is more economic than VCC in maintenance, by replacing the compressor with pump for work addition process.
- the required refrigerant mass flow rate is much higher than VCC for the same evaporator capacity however this increasing in mass flow rate will not affecting on the total volume of the cycle, that because the density of the liquid stat is much higher than density at vapor stat, so the volume of the pump will not increasing as compressors at higher refrigerant mass flow rate, in addition the decreasingin condenser and evaporator effect (heat transfer inKW/Kg of refrigerant mass) will balance the increasing in refrigerant mass flow rate in LCC, so the totalsurface area of the condenser and evaporator in LCC will be close to the VCC, finally the volume of the LCC is approximately as VCC for the same capacity.
- the positive displacement pump could be more suitable for the higher compression ratio comparing to the required mass flow rate.
- the nozzle could be a part of the pump casing, or installed directly after the pump to insure adiabatic compression and expansion.
- a pressurized tank shall installed at the pump suction, also the tank for keeping the condenser, and the suction line at a constant pressure.
- Fig-2 showing the COP levels for Carnot, LCC, and VCC
- Fig-6 showing a schematic drawing for the diversion-conversion diffuser/nozzle
- Ah is Isentropic enthalpy difference
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Abstract
Liquid compression cycle (LCC) is one of the cycles, that can be applied in the refrigeration applications, this cycle has achieved the coefficient of performance of the reversed Carnot cycle, unlike the vapor compression cycle, where a clear deviation from the reversed Carnot cycle is appeared in its ideal case, these deviations from the rev- Carnot cycle have been solved in the Liquid Compression Cycle (LCC) to achieve a thermal efficiency more than the Vapor Compression Cycle (VCC) efficiency.
Description
The Ideal Liquid Compression Refrigeration Cycle
• Technical field
The present invention is directed to the mechanical power engineering for refrigeration and heat pumps.
• Background art
Refrigeration cycles transfer thermal energy from a region of low temperature to one of higher temperature, the reversed Carnot cycle is the perfect model for the refrigeration cycle operating between two fixed temperatures, the most ideal cycle, which has the maximum thermal efficiency, and maximum coefficient of performance, and it serves as a standard against which actual refrigerator cycles can be compared, reversed Camot cycle consist of 4 process, 2 isentropic process for expansion and compression, and 2 isothermal process for heat rejec on and heat absorption.
The ideal Vapor compression cycle, is a cycle built on the principals of the reversed Carnot, but this cycle is deviate from the rev-Carnot for the following reasons:
1 - The refrigerant shall enter the compressor at the vapor phase, for the compressor operation.
2 - Throttling valve is used in expansion process (constant enthalpy process) 3 - The heat rejection and absorption at a constant pressure process, for more practicality.
• Summary of invention
The intent of this invention is to prove a new ideal refrigerator cycle (the Liquid compression cycle) which has a coefficient of performance higher than the Vapor compression cycle.
• Technical problems
The coefficient of performance for the ideal Vapor compression cycle is lower than the reversed Carnot cycle due to the deviation of the ideal process of the VCC than the rev-Carnot, this means that the ideal VCC will consume more electric power than the rev. Carnot at the same refrigerant heat load.
Moreover, all issues related to the compressors in the actual VCC, for example the maintenance, lubrication system, and its higher initial cost,..etc.
• Problems solution
The liquid refrigerant pump in the liquid compression cycle is acting the same function of the compressor in the VCC to solve all the above problems.
• Disclosure of invention
Liquid compression cycle (LCC)is one of the cycles, that can be applied in the refrigeration applications, this cycle has achieved the performance of the reversed Carnot cycle, unlike the vapor compression cycle, where a clear deviation from the reversed Carnot cycle is appeared in its ideal case, the deviation is occur due to the compression process where the refrigerant has to be compressed to a temperature higher than the condensing temperature, and the constant enthalpy process in the expansion valve, where energy loss has occurred due to the irreversibility, these deviations from Carnot cycle have been solved in the Liquid Compression Cycle (LCC)to achieve a thermal efficiency more than the Vapor Compression Cycle (VCC) efficiency, and we will prove that later.
Liquid compression cycle consists of 5 processes, 3 isentropic processes, one isothermal process, and one isobaric process, the cycle (T-H) and (T-S) diagrams are shown in fig-1.
Process (1-2) isentropic compression in a liquid pump
Process (2-3) isentropic expansion in a nozzle
Process (3-4) isothermal heat absorption in an evaporator
Process (4-5) isentropic compression in a diffuser
Process (5-1) isobaric heat rejection in a condenser
Liquid compression cycle is working between 3 levels of pressure, the refrigerant enter the pump at state 1 as a saturated liquid and compressed from the condenser pressure to a higher level pressure, then the refrigerant enters the expansion nozzle to reach the evaporator pressure, during this expansion process the refrigerant lose a lot of internal energy as well as the pressure is decreasing, these amount of energy is converted to kinetic energy at state 3, then the refrigerant is absorbing heat during the isothermal process in the evaporator to reach state 4 in a 2 phase region, then the pressure is regained in the diffuser by converting a part of the kinetic energy again to enthalpy, the refrigerant is isentropic compressed to the condenser pressure at state 5, then the heat is rejected to the ambient at constant pressure to enter the pump again at state 1, g-4 is showing a schema c diagram for the cycle main components.
• Example
The following example is showing how the Liquid compression cycle has achieved the performance of the reversed Carnot cycle comparing with the Vapor compression cycle at the same levels of condenser and evaporator pressure.
As shown in fig-2 a comparison between Carnot, LCC, and VCC according to the COP levels
Assume refrigerant 134a in the Liquid compression cycle is working between the condenser pressure Pi= 1.2 Mpa, and the evaporator pressure P3= 0.36 Mpa, with refrigerant effect 14 kJ/(kg refrigerant), now, we can describe and calculate the properties at each state.
@state 1 saturated liquid phase, Pi= 1.2 Mpa , Ti = 46°C, hi = 117.77 kj/kg, si = 0.424 kj/kg.K, vi = 0.00089 m3/kg.
(gstate 2: sub-cooled phase, P2 shall be calculated by applying the energy equation on the total cycle, as follows : wp = qCo - qev = (Ti - T3) As,
so,As = 14 / (5.8+273) = 0.05 kj/kgK
hence, wp = (46 - 5.8)x0.05 = 2.01kj/kg but, wp = Vi(P2 - P3) , a = (2.01/0.00089) + 1200 = 3458 = 3.46Mpa
, for isentropic compression s2= Si = 0.424 kj/kg.K, calculating h2 = Vi(P2- P3) + hi = (0.0089x(3.46-1.2)xl000) + 117.77 = 119.8 kj/kg.
@state 3 : P3 = 0.36 Mpa, for isentropic expansion s3 = s2 = 0.424 kj/kg.K, T3= 5.8°C, x3= 0.275, calcula ng h3 = hf + x3hfg = 59.72 + (0.275 x 194.08) = 113.1 kj/kg.
@state 4 :P4=P3=0.36 Mpa, calculating s4 = As + s3 = 0.424 + 0.05 = 0.474 kj/kg.K, x4 = 0.347, calcula ng h4 = hf + x4hfg = 59.72 + (0.347 x 194.08) = 127.04kj/ kg
@state 5 :Ps=Pi=1.2 Mpa, for isentropic compression s5 = s4 = 0.474 kj/kg.K, x5 = 0.1015, calculating h5 = hf + x5hfg = 117.77 + (0.102 x 156.1) = 133.61kj/ kg
By comparing these results with the Vapor Compression cycle (VCC) at the same evaporator and condenser pressure as shown in fig-3 :
Estate 1 : Pi = 1.2 Mpa,
117.77 kj/kg
(5) state 2: at throttling process, h2= hi = 117.77 kj/kg
Estate 3: P3 = 0.36 M pa , h3 = 253.81 kj/kg, s3 = 0.9283 kj/kg
(gstate 4: for isentropic compression s4 =s3 = 0.928 kj/kg. K, T4 = 50°C, P4 = 1.2 Mpa, h4 = 278.27 kj/kg
A- The Coefficient of performance calculation (COP):
COPcarnot = T3 / ( T1- T3) = 278.8 / (46 - 5.8 ) = 7 COPLCC = qev/ Wp = 14 / 2 = 7
COPvcc = qev/ wc = (h3 - h2) / (h4 - hi) = 136.04/24.46 = 5.56
B- Special configuration of the Nozzle and Diffuser devices:
In the theoretical study of the liquid compression cycle,a special considerations into nozzle and diffuser shall be considered : i. Diffuser inlet a nd outlet velocities
Defining the relation between the in let and outlet velocities by applying the energy balancing equation on the diffuser, h4 + (V4 2/2) = h5 + (V52/2) ,
(V4 2/2) - (V5 2/2) = AhD
Dividing the two terms by (V 2/2)
V/ νΔ
Defining the relation between the inlet and outlet velocities by a pplying the mass balancing equation on the diffuser,
A . VS A,. V
v5 v4
From equation (la) a nd (2a)
By subs tu ng in equa on 3, where,
v4= 0.0202 m3/kg, and vs= 0.0025 m3/kg (From the previous example)
1 - 0.0153 ^ But from the above relation, we found that;
Hence,
Where, AhD = hs - h4 = 133.61 - 127.04 = 6.57 Kj/kg (From the previous example) Then,
1 ¾ 3.63 m/s
It is difficult to find the exact value for the exit velocity V5, because, it is directly depend on the area ratio A4/A5, which we neglect its value in equation (4a) to find the approximate value of the entering velocity V , so, we go to a probable solution.
When, 1 > (A /As) > 0 then, from equa on (2a) 0.45 > (Vs) > 0
But these values of Vs is too small to be considered as a condenser coil velocities or a pump suction velocities, this will lead to slightly increasing the diffuser outlet velocity, and for solving this problem, the process shall be divided in two sections, as shown in the schematic diagram for the diversion - conversion diffuser, (Fig. no. 6)
Sec on 1, is a diversion diffuser for converting the most amount of kinetic energy to enthalpy energy, and sec on 2 is a conversion nozzle for regaining a small amount of
kinetic energy for increasing condenser and pump inlet velocity as shown in the T - H diagram,(Fig. no. 5)
Assume point C is the intermediate state between sec on 1, and sec on 2, and the refrigerant will reach this stat at 56 °C, hence we can define the point C from the previous example data.
Estate C :Tr = 56 °C, sc = 0.474 KJ/Kg.K , xc = 0.0087, hc = 134.16 KJ/kg, vc = 0.00103 m3/kg
By applyingequation4a for sec on 1, as the ow entering the sec on at stat 4, and leaving at stat C
.04 = 7.12 KJ/kg
From the above result, the probable solution for the exit velocity Vc
When, 1 > (A4/Ac) > 0 then, from equa on (2a) 0.19> (VC) > 0
By applying equation 4a for section 2, as the flow entering the section at stat C and leaving at stat 5
Where,AhD2 = hc - h5 = 134.16 - 133.61 = 0.55 Kj/kg
From the above result, the probable solution for the inlet velocity Vc
When, 1 > (Ac/A5) > 0 then, from equa on (2a) 0.43> (Vc) > 0
From sec on 1 and sec on 2, it is clearthat the common interval value of the exit velocity Vc
0.19 > (VC) > 0 then, from equa on (2a) 0.44 > (Ac/A5) > 0 Note: When, 1 > (A4/Ac) > 0 this means 0.44 > (Ac/A5) >0 , and leads to A4/A5 = 0.44 i. Nozzle inlet and outlet velocities
Defining the relation between the inlet and outlet velocities by applying the energy balancing equation on the diffuser, h2 + (V2 2/2) = h3 + (V3 2/2) ,
(V3 2/2) - (V2 2/2) = AhN
Dividing the two terms by (V3 2/2)
V2 _ 2AhN
V2 V2
2AhN
{lb)
Defining the relation between the inlet and outlet velocities by applying the mass balancing equation on the diffuser,
A3. V3 = A2. V2
v3 v2
From equation (lb) and (2b)
By subs tu ng in equa on 3b, Where, v3= 0.016 m3Ag, and v2= 0.00089 m3/kg (From the previous example)
From the above equation, we find that;
A2
1 - 0.003 - ~ 1
Ai
1
Hence,
Where, AhN = h2 - h3 = 119.8 - 113.1 = 13.4 Kj'/kg Then, V3 « 3.67 m/s
It is difficu lt to find the exact value for the inlet velocity V2, beca use, it is directly depend on the a rea ratio A3/A2, which we neglect its value in equa on (4b) to find the approximate value of the exit velocity V3, so, we go to a probable solution.
When, 1 > (A3/A2) > 0 then, from equa on (2b) . 0.21> (V2) > 0
But, V2 is representing the pump discharge velocity, so the above value is too small and not practical.
Assume that the pump neither adding nor extracting kinetic energy from the fluid, so, i = V2 = 1.05 m/s
To apply this value at the pum p discharge, we shal l split the Nozzle into two sections, diversion section, and conversion section.
Sec on 1, is a diversion di user for conver ng small amount of the kine c energy to enthalpy energy till the velocity at the diffuser reach the required value of the nozzle inlet, then the ow will entering sec on 2, where a conversion nozzle is regain most amount of the kinetic energy, till the velocity at the nozzle outlet is reaching the value 3.67 m/s. as shown in the T - H diagram, (Fig. no. 5).
Assume point E is the intermediate state between sec on 1, and sec on 2, and the refrigerant wil l reach this stat at PE = 4 Mpa, hence we can de ne the point E from the previous example data.
Estate E : vE=v2 = i= 0.00089 m3/kg, sE = 0.424 KJ/Kg.K, fall in the sub-cooled liquid region.
By applying equa on 4b for sec on 1, as the flow entering the section at stat 2, and leaving at stat E
ΔΙΊΝΙ =hE- h2 = vE (PE - P2) =0.00089(4000 - 3460) = 0.48 KJ/kg
But, V2 = Vi = 1.05 m/s
Apply the energy ba lancing between 2 and E
Then,
(VE 2/2) = (V2 2/2) - ΔηΝ1= 0.55 - 0.48 = 0.07 VE = 0.374 m/s
Applying equa on 2b between 2 and E
By applying equa on 4b for sec on 2, as the flow entering the section at stat E, and leaving at stat 3
Where,
AhN2 = hE-h3=h2 + ΔΙΊΝΙ - h3 = (119.8 +0.48 ) - (113.1) = 7.18 Kj/kg
From equa on 4b
V3 « J2AhN2 = 3.78 m/s
It is difficult to find the exact value for the inlet velocity VE, because, it is directly depend on the area ratio A3/AE, which we neglect its value in equa on (4b) to nd the approximate value of the exit velocity V3, so, we go to a probable solution.
When, 1 > (A3/A2) > 0 then, from equa on (2b) 0.21 > (VE) > 0
Note: From the mass balancing on sec on 2, we found that 0.21 > (VE) > 0 , however, in sec on 1, VE = 0.377 m/s, this means that the assumption value of the pressure at E (PE) shall be slightly increased, to insure that the intermediate velocity VE will fall between 0.21 > (VE) > 0
C- General configuration on the actual Liquid compression cycle :
1 - As shown in the previous example the higher pressure level is calculated according to the minimum potential work needed for the reversible Liquid compression cycle, in the actual cycle, that pressure shall be increased to overcome the irreversibility in the cycle.
2 -The expansion process occur in the nozzle will be adiabatic irreversible process, and,
(Actual Kinetic energy at exit)
' Isentropic Kinetic energy at exit)
3 -The compression process occur in the diffuser will be adiabatic irreversible process,
Δηί5
_ Isentropic Kinetic energy at exit) I _ ¼s
~~ I (Actual Kinetic energy at exit) ~ va 2 ct.
4 -The compression process occur in the pump will be adiabatic irreversible process,
5 - A pressure drop shall occur in evaporator and condenser coil, the same as the actual Vapor compression cycle.
• Advantages of the Liquid Compression Cycle on the Vapor Compression Cycle :
1 - The coefficient of performance of LCC is higher than VCC.
2 - If the refrigerant exit the condenser in the sub-cooled region, or state 1 is fall in the sub-cooled region, the COP of the LCC will slightly raised above the reversed Carnot cycle, the limitation for this raise depending on the minimum temperature approach between the refrigerant and the ambient or the cooling medium.
3 - The work addition process is occur in the liquid phase, thus the actual process will be close to the isentropic process, unlike the VCC, the work addition process in the superheat region, where the entropy deviation increased, and the irreversibility is increased in the actual cycle.
4 - The constant enthalpy process in the expansion valve, increasing the
irreversibility in the actual VCC cycle.
5 -The low refrigerant velocity for the vapor line and condenser coil, will decreasing the friction loss in pipes, and hence the irreversibility (or energy loss) will decrease in LCC comparing with VCC.
6 - The lubrication system challenges in the VCC are not exist in the LCC by separating the lubricant from the refrigerant path.
7 - The LCC is more economic than VCC in maintenance, by replacing the compressor with pump for work addition process.
8 - The initial cost of LCC is lower than VCC in the reason of using pump instead of the compressor.
• Disadvantages of the Liquid Compression Cycle on the Vapor Compression Cycle
1 - The required refrigerant mass flow rate is much higher than VCC for the same evaporator capacity however this increasing in mass flow rate will not affecting on the total volume of the cycle, that because the density of the liquid stat is much
higher than density at vapor stat, so the volume of the pump will not increasing as compressors at higher refrigerant mass flow rate, in addition the decreasingin condenser and evaporator effect (heat transfer inKW/Kg of refrigerant mass) will balance the increasing in refrigerant mass flow rate in LCC, so the totalsurface area of the condenser and evaporator in LCC will be close to the VCC, finally the volume of the LCC is approximately as VCC for the same capacity.
2 - The high refrigerant velocity for the liquid line and evaporator coil, will increasing the friction loss in pipes, and hence the irreversibility (or energy loss) will increase in LCC comparing with VCC.
• General recommendation on the Liquid compression cycle :
1 - The positive displacement pump could be more suitable for the higher compression ratio comparing to the required mass flow rate.
2 - Installing a refrigerant distributor before the evaporator and condenser coils to divide the mass flow rate on a multiple paths, to increase the heat transfer area and decreasing the refrigerant paths, also the nozzle could be a part of the pump casing, or installed directly after the pump to insure adiabatic compression and expansion.
3- For preventing cavitations at the centrifugal pump suction line a pressurized tank shall installed at the pump suction, also the tank for keeping the condenser, and the suction line at a constant pressure.
• Brief description of the drawings
Fig-1 showing the LCC on T-S and T-H diagram
Fig-2 showing the COP levels for Carnot, LCC, and VCC
Fig-3 showing the VCC on T-S and T-H diagrams
Fig-4 showing a simple schematic diagram for the main components
Fig-5 showing the diversion-conversion diffuser/nozzle on T-H diagram
Fig-6 showing a schematic drawing for the diversion-conversion diffuser/nozzle
• Legend:
LCC Liquid Compression Refrigeration Cycle VCC Vapor Compression Refrigeration Cycle T Temperature P Pressure
h enthalpy per unit of refrigerant mass
s entropy per unit of refrigerant mass wp Mechanical pump work per unit of refrigerant mass
wc Mechanical compressor work per unit of refrigerant mass
qco Heat rejected from condenser per unit of refrigerant mass
qev Heat absorbed to evaporator per unit of refrigerant mass
COP Coefficient Of Performance
x Mass quality, the ratio of the vapor mass to the total mass of the mixture.
AhN Total Enthalpy difference in the diversion-conversion nozzle
AhNi Enthalpy difference in the diversion section of the diversion-conversion nozzle
ΔηΝ2 Enthalpy difference in the conversion section of the diversion-conversion nozzle
AhD Total Enthalpy difference in the diversion-conversion diffuserAhD= AhD1-
AhDi Enthalpy difference in the diversion section of the diversion-conversion diffuser
AhD2 Enthalpy difference in the conversion section of the diversion-conversion diffuser
P Pump.
N Nozzle.
D Diffuser.
CO Condenser coil.
EV Evaporator coil.
Ahact Actual enthalpy difference
Ahis Isentropic enthalpy difference
Actual Kinetic energy
Vis 2/2 Isentropic Kinetic energy ηί5 N Isentropic efficiency of the nozzle ηί5 θ Isentropic efficiency of the diffuser
Claims
Claims
Liquid compression refrigeration cycle (LCC) is a new operating theory for refrigeration and heat pump applications, the cycle is consisting of 5 processes, 3 isentropic processes, one isothermal process, and one isobaric process, the cycle (T- H) and (T-S) diagrams are shown in fig-1.
Process (1-2) isentropic compression in a liquid pump
Process (2-3) isentropic expansion in a nozzle
Process (3-4) isothermal heat absorption in an evaporator coil
Process (4-5) isentropic compression in a di user
Process (5-1) isobaric heat rejec on in a condenser coil
Liquid compression cycle is working between 3 levels of pressure, the refrigerant enter the pump at state 1 as a saturated liquid and compressed from the condenser pressure to a higher level pressure, then the refrigerant enters the expansion nozzle to reach the evaporator pressure, during this expansion process the refrigerant lose a lot of internal energy as well as the pressure is decreasing, these amount of energy is converted to kine c energy at state 3, then the refrigerant is absorbing heat during the isothermal process in the evaporator to reach state 4 in a 2 phase region, then the pressure is regained in the diffuser by converting a part of the kinetic energy again to enthalpy, the refrigerant is isentropic compressed to the condenser pressure at state 5, then the heat is rejected to the ambient at constant pressure to enter the pump again at state 1, g-4 is showing a schema c diagram for the cycle main components.
A special configuration for the Nozzle and Diffuser devices
Diversion - conversion diffuser (Fig. no. 6) is a diffuser divided in to two sections:
Sec on 1, is a diversion diffuser for converting the most amount of kinetic energy to enthalpy energy, and sec on 2 is conversion nozzle for regaining a small amount of kinetic energy for increasing condenser and pump inlet velocity as shown in the T - H diagram,(Fig. no. 5)
Diversion - conversion nozzle (Fig. no. 6) is a nozzle divided in to two sections:
Sec on 1, is a diversion di user for conver ng small amount of the kine c energy to enthalpy energy till the velocity at the diffuser reach the required value at the nozzle inlet, then the ow will entering sec on 2, where a conversion nozzle is regain most amount of the kinetic energy, till the velocity at the nozzle outlet is reaching the value 3.67 m/s. as shown in the T - H diagram, (Fig. no. 5).
Priority Applications (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| PCT/EG2015/000005 WO2016134731A2 (en) | 2015-02-25 | 2015-02-25 | The ideal liquid compression refrigeration cycle |
| US15/686,739 US20180045440A1 (en) | 2015-02-25 | 2017-08-25 | Ideal Liquid Compression Refrigeration Cycle |
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| PCT/EG2015/000005 WO2016134731A2 (en) | 2015-02-25 | 2015-02-25 | The ideal liquid compression refrigeration cycle |
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| US15/686,739 Continuation-In-Part US20180045440A1 (en) | 2015-02-25 | 2017-08-25 | Ideal Liquid Compression Refrigeration Cycle |
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| WO2016134731A2 true WO2016134731A2 (en) | 2016-09-01 |
| WO2016134731A3 WO2016134731A3 (en) | 2017-06-01 |
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| US (1) | US20180045440A1 (en) |
| WO (1) | WO2016134731A2 (en) |
Cited By (2)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| WO2016134731A3 (en) * | 2015-02-25 | 2017-06-01 | Hossain Khaled Mohammed | The ideal liquid compression refrigeration cycle |
| CN109447483A (en) * | 2018-11-01 | 2019-03-08 | 国电科学技术研究院有限公司 | A kind of calculation method of low-level (stack-gas) economizer to Specific Heat Consumption For Steam Turbine Unit influence amount |
Families Citing this family (2)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US10808646B2 (en) | 2019-01-09 | 2020-10-20 | Haier Us Appliance Solutions, Inc. | Cooled piston and cylinder for compressors and engines |
| CN111721019B (en) * | 2019-01-09 | 2023-03-24 | 李华玉 | First-class thermally-driven compression heat pump |
Family Cites Families (9)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US5343711A (en) * | 1993-01-04 | 1994-09-06 | Virginia Tech Intellectual Properties, Inc. | Method of reducing flow metastability in an ejector nozzle |
| US5819554A (en) * | 1995-05-31 | 1998-10-13 | Refrigeration Development Company | Rotating vane compressor with energy recovery section, operating on a cycle approximating the ideal reversed Carnot cycle |
| ES2235681T3 (en) * | 2003-07-18 | 2010-08-31 | Star Refrigeration Ltd. | IMPROVED TRANSCRITICAL COOLING CYCLE. |
| CN1304797C (en) * | 2005-01-10 | 2007-03-14 | 深圳清华大学研究院 | Steam compression type refrigeration cycle device of approximate ideal inverse Carnot cycle efficiency |
| US20060242992A1 (en) * | 2005-05-02 | 2006-11-02 | Mark Nicodemus | Thermodynamic apparatus and methods |
| JP4737001B2 (en) * | 2006-01-13 | 2011-07-27 | 株式会社デンソー | Ejector refrigeration cycle |
| EP1999415B1 (en) * | 2006-03-27 | 2011-07-27 | Mayekawa Mfg Co.Ltd. | Refrigerating apparatus with vapor compression refrigerating cycle and control method therefor |
| US20110225990A1 (en) * | 2010-03-19 | 2011-09-22 | Air Generate Inc | Efficient heat pump |
| WO2016134731A2 (en) * | 2015-02-25 | 2016-09-01 | Hossain Khaled Mohammed | The ideal liquid compression refrigeration cycle |
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- 2015-02-25 WO PCT/EG2015/000005 patent/WO2016134731A2/en active Application Filing
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Cited By (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| WO2016134731A3 (en) * | 2015-02-25 | 2017-06-01 | Hossain Khaled Mohammed | The ideal liquid compression refrigeration cycle |
| CN109447483A (en) * | 2018-11-01 | 2019-03-08 | 国电科学技术研究院有限公司 | A kind of calculation method of low-level (stack-gas) economizer to Specific Heat Consumption For Steam Turbine Unit influence amount |
| CN109447483B (en) * | 2018-11-01 | 2021-11-12 | 国电科学技术研究院有限公司 | Method for calculating influence quantity of low-temperature economizer on heat consumption rate of steam turbine unit |
Also Published As
| Publication number | Publication date |
|---|---|
| US20180045440A1 (en) | 2018-02-15 |
| WO2016134731A3 (en) | 2017-06-01 |
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